Control device for spark-ignition engine

ABSTRACT

The invention is intended to provide improved emission-cleaning performance by use of a three-way catalyst alone, without the need for a lean NOx catalyst, while ensuring a fuel economy improvement effect of lean burn operation. A multicylinder spark-ignition engine is constructed such that, in a pair of preceding and following cylinders whose exhaust and intake strokes overlap each other, burned gas discharged from the preceding cylinder ( 2 A,  2 D) which is currently in the exhaust stroke is introduced directly into the following cylinder ( 2 B,  2 C) which is currently in the intake stroke through an intercylinder gas channel ( 22 ) and gas discharged from only the following cylinder ( 2 B,  2 C) is led to an exhaust passage ( 20 ) provided with a three-way catalyst ( 24 ) in a low-load, low-speed operating range. Fuel supply to the individual cylinders is controlled in such a manner that combustion in the preceding cylinder ( 2 A,  2 D) is made under lean mixture conditions at an air-fuel ratio larger than the stoichiometric air-fuel ratio by a specific amount and combustion in the following cylinder ( 2 B,  2 C) is made under conditions of the stoichiometric air-fuel ratio created by supplying fuel to the burned gas introduced from the preceding cylinder ( 2 A,  2 D).

TECHNICAL FIELD

This invention relates to a control device for a spark-ignition engine.More particularly, the invention pertains to a device for controllingcombustion state in individual cylinders to improve fuel economy andcharacteristics of emissions in a multicylinder engine.

BACKGROUND ART

There exists a conventionally known technique for achieving animprovement in fuel economy by performing combustion in a state of“lean” air-fuel ratio in which air-fuel mixture in individual cylindersof a spark-ignition engine is burnt at an air-fuel ratio larger than thestoichiometric air-fuel ratio. One example of this kind is shown inJapanese Unexamined Patent Publication No. H10-274085, which employsfuel injectors for injecting fuel directly into combustion chambers tocause stratified charge combustion by injecting fuel during acompression stroke in a low-speed, low-load range, for example, tothereby accomplish extremely lean mixture combustion.

In this kind of engine, it is impossible to achieve sufficientemission-cleaning performance with respect to nitrogen oxides (NOx)under lean burn operating conditions by using an ordinary three-waycatalyst alone, which is a catalyst having high performance to converthydrocarbons (HC), carbon monoxide (CO) and NOx at about thestoichiometric air-fuel ratio, as an emission-cleaning converter.Therefore, as shown in the aforementioned Publication, the engine isprovided with a lean NOx catalyst which adsorbs NOx in an oxygen-richatmosphere and releases and reduces NOx in an atmosphere where oxygenconcentration has decreased. If the amount of NOx adsorbed by the leanNOx catalyst has increased under the lean burn operating conditions whenthe lean NOx catalyst of this kind is being used, the fuel is injectednot only for primary combustion but an additional amount of fuel isinjected during an expansion stroke to decrease the air-fuel ratio andgenerate CO for accelerating release and reduction of NOx as shown inthe aforementioned Publication, for example.

The aforementioned engine which performs conventional lean burnoperation requires the lean NOx catalyst to provide NOx-convertingperformance during the lean burn operating conditions. This type ofengine also requires the three-way catalyst for cleaning emissions insuch engine operating ranges as a high-load range in which the engine isoperated at the stoichiometric air-fuel ratio. The lean NOx catalystprovided along with the three-way catalyst needs to have a relativelylarge capacity to provide a capability to adsorb a certain amount of NOxand is expensive as compared to the three-way catalyst, so that theprovision of this lean NOx catalyst is disadvantageous from theviewpoint of product cost.

In addition, it is necessary to temporarily decrease the air-fuel ratioby feeding additional amounts of fuel to accelerate release andreduction of NOx at specific intervals of time when the amount of NOxadsorbed increases as stated above in order to maintain the convertingperformance of the lean NOx catalyst. This would jeopardize fuel economyimprovement effect offered by lean burn operation.

Furthermore, the lean NOx catalyst is susceptible to poisoning bysulfurization when used fuel contains high sulfur content. The lean NOxcatalyst should therefore be subjected to regeneration treatment, suchas catalyst heating and feeding of a reducing agent, to prevent thissulfur-poisoning problem. This regeneration treatment of the lean NOxcatalyst is likely to cause a reduction in the fuel economy improvementeffect and deterioration of its durability.

The invention has been made in consideration of the aforementionedproblems of the prior art. Accordingly, it is an object of the inventionto provide a control device of a spark-ignition engine capable ofproviding improved emission-cleaning performance by use of a three-waycatalyst alone, without the need for a lean NOx catalyst, while ensuringa fuel economy improvement effect of lean burn operation.

DISCLOSURE OF THE INVENTION

A control device of the invention is for use in a multicylinderspark-ignition engine of which individual cylinders go throughsuccessive cycles of intake, compression, expansion and exhaust strokeswith specific phase delays, in which fresh air and gas flow paths areconnected to form a two-cylinder interconnect configuration at least ina low-load, low-speed operating range such that, in a pair of precedingand following cylinders whose exhaust and intake strokes overlap eachother, burned gas discharged from the preceding cylinder which iscurrently in the exhaust stroke is introduced directly into thefollowing cylinder which is currently in the intake stroke through anintercylinder gas channel and gas discharged from only the followingcylinder is led to an exhaust passage provided with a three-waycatalyst. The control device comprises a controller for controlling fuelsupply to the individual cylinders in such a manner that combustion inthe preceding cylinder is made under lean mixture conditions at anair-fuel ratio larger than the stoichiometric air-fuel ratio by aspecific amount, fuel for the following cylinder is supplied to theburned gas generated by combustion in the preceding cylinder, andcombustion in the following cylinder is made under conditions of thestoichiometric air-fuel ratio when the engine is in the two-cylinderinterconnect configuration.

In this construction, combustion in the preceding cylinder is made at a“lean” air-fuel ratio at least in the low-load, low-speed operatingrange so that a significant fuel economy improvement effect is achieveddue to an increase in thermal efficiency and a reduction in pumping lossand the amount of NOx generated in the preceding cylinder is kept to arelatively low level. On the other hand, since the burned gas isintroduced from the preceding cylinder into the following cylinder, acondition equivalent to what would occur when a great deal of exhaustgas is introduced by exhaust gas recirculation (EGR) is created in thefollowing cylinder. As a result, pumping loss is reduced and NOxemission is sufficiently decreased in the following cylinder. Inaddition, since combustion in the following cylinder is made underconditions of the stoichiometric air-fuel ratio created by supplying thefuel to the burned gas, sufficient emission-cleaning performance isachieved by the three-way catalyst alone and the provision of a lean NOxcatalyst becomes unnecessary.

While high-temperature gas discharged from the preceding cylinder passesthrough the intercylinder gas channel, gas temperature is regulated bymoderate heat dissipation along its length. Because this gas in whichthe burned gas and excess air are mixed and dispersed uniformly isintroduced into the following cylinder, an ideal condition is createdfor introduction of a great deal of EGR gas. In addition, since the fuelis injected into the gas at a relatively high temperature, evaporationof the fuel is accelerated and combustion in the following cylinder isperformed in a desirable fashion.

In the control device for the spark-ignition engine of the invention,two cylinders may constitute the aforesaid pair of the preceding andfollowing cylinders if the exhaust stroke of one cylinder perfectlycoincides in timing with the intake stroke of the other cylinder, or ifthe exhaust stroke of one cylinder precedes and coincides in part withthe intake stroke of the other cylinder. If the preceding and followingcylinders meet these conditions, the burned gas discharged from thepreceding cylinder is smoothly introduced into the following cylinderthrough the intercylinder gas channel and pumping loss is effectivelyreduced.

Preferably, the control device further comprises a flow path switcherfor switching the fresh air and gas flow paths to form an independentcylinder configuration in a high-load, high-speed operating range, inwhich intake ports and exhaust ports of the individual cylinders workindependently of one another such that fresh air is introduced throughan intake passage into the intake ports of the individual cylinders andexhaust gas discharged through the exhaust ports of the individualcylinders is led to the exhaust passage, wherein the aforementionedcontroller makes the air-fuel ratio in the individual cylinders equal toor smaller than the stoichiometric air-fuel ratio in the high-load,high-speed operating range.

This construction serves to improve fuel economy and characteristics ofemissions in the low-load, low-speed operating range and to ensureengine output performance in the high-load, high-speed operating range.

It is preferable in the above construction that the preceding cylinderbe provided with an intake port connected to the intake passage, a firstexhaust port connected to the exhaust passage and a second exhaust portconnected to the intercylinder gas channel, the following cylinder beprovided with a first intake port connected to the intake passage, asecond intake port connected to the intercylinder gas channel and anexhaust port connected to the exhaust passage, and the flow pathswitcher include a valve stop mechanism which individually switchesfirst and second exhaust valves for opening and closing the first andsecond exhaust ports of the preceding cylinder as well as first andsecond intake valves for opening and closing the first and second intakeports of the following cylinder between activated and deactivatedstates, and a valve stop mechanism controller which sets the firstexhaust valve and the first intake valve to the deactivated state andthe second exhaust valve and the second intake valve to the activatedstate in the low-speed, low-load operating range, and sets the firstexhaust valve and the first intake valve to the activated state and thesecond exhaust valve and the second intake valve to the deactivatedstate in the high-load, high-speed operating range.

This construction makes it possible to easily switch the fresh air andgas flow paths in a manner suited for the low-speed, low-load operatingrange and the high-load, high-speed operating range by controlling thevalve stop mechanism.

Preferably, the engine is made switchable between special operation modein which combustion is made in the two-cylinder interconnectconfiguration and normal operation mode in which combustion is made withthe intake ports and the exhaust ports of the individual cylindersworking independently of one another according to engine operatingconditions, and the control device further comprises an intake airpulsation detector for detecting intake air pulsations, wherein thecontrol device judges at switching of the engine operation mode whetherthe fresh air and gas flow paths have been switched by the flow pathswitcher with reference to a sensing signal output from the intake airpulsation detector and performs air-fuel ratio control operationcorresponding to the operation mode selected after switching of the flowpaths following a point in time when the switching of the flow paths hasbeen detected.

In this construction, the control device performs the air-fuel ratiocontrol operation corresponding to the newly selected operation modeupon confirming the completion of the switching of the flow paths by theflow path switcher when the engine has been switched between the specialoperation mode and the normal operation mode.

In a case where the operation mode is switched according to the engineoperating conditions as stated above, the control device shouldpreferably judge that the switching of the flow paths has been completedat a point in time when a sudden change in the period of intake airpulsations is verified with reference to a sensing signal output from anintake air quantity detector for detecting the amount of intake air.

This construction makes it possible to exactly judge whether theswitching of the flow paths has actually been done by detecting thechange in the period of intake air pulsations occurring at the switchingof the intake and exhaust flow paths with reference to the sensingsignal output from the intake air quantity detector.

In checking the sudden change in the period of intake air pulsations inthe aforementioned manner, it is preferable to judge that the switchingof the flow paths from the two-cylinder interconnect configuration tothe independent cylinder configuration has been done when it is verifiedthat the period of intake air pulsations has become shorter.

According to this arrangement, the period of intake air pulsationsbecomes shorter as a result of an increase in the number of intake airpulsations per unit time when the intake and exhaust flow paths areswitched from the two-cylinder interconnect configuration to theindependent cylinder configuration creating a condition in which freshair is introduced into the individual cylinders. It is possible toexactly judge that the switching from the two-cylinder interconnectconfiguration to the independent cylinder configuration has been done bythe flow path switcher based on this change in the period of intake airpulsations.

It is preferable to change the intake and exhaust flow paths by varyingthe amount of valve lift determined by a valve actuating mechanismprovided to each cylinder.

According to this arrangement, switching of the operation mode betweenthe special operation mode and the normal operation mode is quickly doneand the air-fuel ratio control operation corresponding to the operationmode selected after the switching of the flow paths is performed afterit has been verified that operation for switching the flow paths byvarying the amount of valve lift determined by the valve actuatingmechanism provided to each cylinder has been completed with reference tothe sensing signal output from the intake air pulsation detector.

In the case where the operation mode is switched according to the engineoperating conditions as stated above, it is preferable to providemultiple pairs of the preceding and following cylinders of which intakeand exhaust strokes overlap with each other and to perform the air-fuelratio control operation corresponding to the operation mode selectedafter the switching of the flow paths in all the pairs of the precedingand following cylinders following a point in time when the switching ofthe flow paths has been first verified in one of the multiple pairs ofthe multiple pairs of the preceding and following cylinders.

According to this arrangement, the air-fuel ratio control operationcorresponding to the operation mode selected after the switching of theflow paths is performed on all the pairs of the preceding and followingcylinders at the point in time when the switching of the flow paths hasbeen verified in one pair of the preceding and following cylinders withreference to sensing signal output from the intake air pulsationdetector while the switching of the flow paths in the multiple pairs ofthe preceding and following cylinders is made in a specific order. As aconsequence, the air-fuel ratio control operation corresponding to thenewly selected operation mode is executed quickly and properly.

In the case where the operation mode is switched according to the engineoperating conditions as stated above, it is preferable to judge that theflow paths have been switched at a point in time when a switching signalhas been output to the flow path switcher following a change in theengine operating conditions and the occurrence of a change in intake airpulsations has been verified with reference to the sensing signal outputfrom the intake air pulsation detector.

This arrangement makes it possible to prevent an incorrect judgment dueto sensing errors of the intake air pulsation detector or noisecontained in its sensing signal and perform the air-fuel ratio controloperation corresponding to the operation mode selected after theswitching of the flow paths after it has been exactly verified that theswitching of the flow paths has been done in accordance with theaforementioned sensing signal.

If the control device is caused to begin preparation for executing theair-fuel ratio control operation after the switching of the engineoperation mode at a point in time when it is verified that the flow pathswitching signal has been output to the flow path switcher, the air-fuelratio control operation corresponding to the operation mode selectedafter the switching of the flow paths is quickly performed after it hasbeen exactly verified that the switching signal has been output to theflow path switcher and the switching of the flow paths has been done asa result of the occurrence of a change in intake air pulsationsaccording to the sensing signal output from the intake air pulsationdetector.

Preferably, the control device of the invention further comprises anexhaust gas concentration detector disposed in the exhaust passageprovided with the three-way catalyst for detecting the stoichiometricair-fuel ratio and an exhaust gas concentration detector disposed in theintercylinder gas channel for detecting a lean mixture state, whereinthe controller feedback-controls the amounts of fuel injected into theindividual cylinders based on values detected by the individual exhaustgas concentration detectors in such a manner that the air-fuel ratio inthe preceding cylinder becomes larger than the stoichiometric air-fuelratio by a specific amount and the air-fuel ratio in the followingcylinder becomes equal to the stoichiometric air-fuel ratio when theengine is in the two-cylinder interconnect configuration.

This construction makes it possible perform such control operation highaccuracy that produces a specific air-fuel ratio in the precedingcylinder and the stoichiometric air-fuel ratio in the followingcylinder.

Preferably, the control device of the invention should be such that theengine has a fuel injector for injecting fuel directly into thepreceding cylinder and the aforementioned controller causes stratifiedcharge combustion to occur in the preceding cylinder by injecting thefuel during its compression stroke from the fuel injector whileproducing a lean mixture state therein when the engine is in thetwo-cylinder interconnect configuration.

In this construction, combustion in the preceding cylinder is made in adesirable fashion by stratification even at a “lean” air-fuel ratio.

If the engine is further provided with a fuel injector for injectingfuel directly into the following cylinder and the controller causesstratified charge combustion to occur in the following cylinder byinjecting at least part of the fuel during its compression stroke whileproducing the stoichiometric air-fuel ratio therein when the engine isin the two-cylinder interconnect configuration, stratified chargecombustion or slightly stratified charge combustion is produced in thefollowing cylinder, so that combustion in the following cylinder is madein a desirable fashion even in a condition equivalent to what wouldoccur when a great deal of exhaust gas is introduced by EGR operation.

Alternatively, the controller may cause uniform charge combustion tooccur in the following cylinder while producing the stoichiometricair-fuel ratio therein when the engine is in the two-cylinderinterconnect configuration. This is effective if it is possible tomaintain ignitability even when the fuel is uniformly dispersed in thefollowing cylinder due to a sufficiently high temperature of the burnedgas introduced from the preceding cylinder into the following cylinder,for example.

The engine may be provided with a fuel injector disposed in an intakepassage of the following cylinder for injecting fuel directly into thefollowing cylinder, the intake passage constituting the intercylindergas channel, and the controller may cause uniform charge combustion tooccur in the following cylinder by injecting the fuel during its intakestroke while producing the stoichiometric air-fuel ratio therein whenthe engine is in the two-cylinder interconnect configuration.

According to this arrangement, heat of exhaust gas introduced from thepreceding cylinder into the following cylinder is moderately dissipated,and the fuel is supplied to a great deal of ideal EGR gas in whichexcess air and burned gas are mixed, so that mixing of the fuel with theEGR gas and evaporation of the fuel are accelerated and combustibilityin the following cylinder is improved with a great deal of EGR gasintroduced thereinto. As a result, emission and fuel economy performanceis improved and the possibility of self-ignition operation of thefollowing cylinder is increased due to improvements in mixability of theburned gas, air and fuel and in evaporability of the fuel even when agreat deal of exhaust gas is introduced by EGR operation.

When the engine is in the two-cylinder interconnect configuration, it ispreferable to make the air-fuel ratio in the preceding cylinderapproximately equal to twice the stoichiometric air-fuel ratio orlarger. This arrangement serves to sufficiently increase fuel economyimprovement effect gained by the “lean” air-fuel ratio, prevent theamount of excess air in the burned gas introduced into the followingcylinder from becoming too small, and ensure combustibility in thefollowing cylinder.

Preferably, the control device of the invention should be such that thecontroller includes a total fuel injection quantity calculator forcalculating the sum of the amounts of fuel to be injected into thepreceding cylinder and the following cylinder based on the amount ofintake air introduced into the preceding cylinder in such a manner thatcombustion in the preceding cylinder is made under the lean mixtureconditions at the air-fuel ratio larger than the stoichiometric air-fuelratio by the specific amount and combustion in the following cylinder ismade under the conditions of the stoichiometric air-fuel ratio when theengine is in the two-cylinder interconnect configuration, a ratio setterfor setting a ratio of the air-fuel ratio for the preceding cylinder tothe air-fuel ratio for the following cylinder according to engineoperating conditions in such a manner that a balance is achieved betweena torque generated by the preceding cylinder and a torque generated bythe following cylinder when the engine is in the two-cylinderinterconnect configuration, and a final fuel injection quantitycalculator for calculating final amounts of fuel to be injected into thepreceding cylinder and the following cylinder based on the ratio of theair-fuel ratios set by the ratio setter and the sum of the amounts offuel to be injected calculated by the total fuel injection quantitycalculator.

In this construction, the amounts of fuel to be injected into thepreceding cylinder and the following cylinder is determined as describedabove and the amounts of fuel so determined are injected such that abalance is achieved between the torques generated by the precedingcylinder and the following cylinder in the special operation mode inwhich combustion is made in the two-cylinder interconnect configuration.Consequently, the occurrence of a difference in the torques generated bythe preceding cylinder and the following cylinder due to a difference inthermal efficiency including pumping loss is avoided or suppressed.

The ratio of the air-fuel ratios set by the ratio setter may bedetermined based on a parameter concerning pumping loss or thermalefficiency of the preceding cylinder and the following cylinder. Theaforementioned ratio of the air-fuel ratios may be determined from dataon experimentally obtained torques generated by the preceding cylinderand the following cylinder or from parameters concerning the pumpingloss and thermal efficiency of the preceding cylinder and the followingcylinder.

Preferably, the aforementioned controller further includes acombustibility judgment section for judging in advance whether the stateof combustion in the preceding cylinder and the following cylinderpredicted based on the ratio of the air-fuel ratios obtained by theratio setter falls in a normally combustible range, wherein the finalfuel injection quantity calculator calculates the final amounts of fuelto be injected based on the ratio of the air-fuel ratios set by theratio setter only if the judgment result of the combustibility judgmentsection is in the affirmative.

In this construction, calculation of the amounts of fuel to be injectedbased on the ratio of the air-fuel ratios determined as described aboveis prohibited so that the occurrence of misfire or abnormal combustion(knocking) can be avoided if it is difficult to make normal combustionin the preceding cylinder or the following cylinder according to thatratio of the air-fuel ratios even when the ratio of the air-fuel ratioswhich could theoretically achieve a torque balance between the precedingcylinder and the following cylinder has been obtained.

If the judgment result of the combustibility judgment section is in thenegative, the final fuel injection quantity calculator should preferablycalculate the final amounts of fuel to be injected into the individualcylinders based on a ratio preset within a range in which normalcombustion can be made in the individual cylinders. This arrangementserves to prevent the occurrence of misfire and abnormal combustion andmake normal combustion in the preceding cylinder and the followingcylinder.

Preferably, the aforementioned controller further includes an ignitioncontroller for selecting whether to produce combustion in the followingcylinder by compressed self-ignition or by forced ignition according tothe engine operating conditions when the engine is in the two-cylinderinterconnect configuration, wherein the ratio setter varies the ratio ofthe air-fuel ratios to be set depending on whether combustion is made bycompressed self-ignition or by forced ignition.

In this construction, the difference in the torques generated by thepreceding cylinder and the following cylinder would increase whencombustion in the following cylinder is made by compressed self-ignitionin the special operation mode, because the thermal efficiency in thefollowing cylinder is improved. Since a ratio of the air-fuel ratiosdifferent from that for a case where compressed self-ignition is notperformed, or a ratio determined taking into account the improvement inthe thermal efficiency in the following cylinder, is set, however, theoccurrence of a difference in the torques generated by the precedingcylinder and the following cylinder is avoided or suppressed as in thecase where compressed self-ignition is not performed.

Preferably, the control device of the invention should be such that theengine is controlled in a manner that makes combustion by compressedself-ignition in the following cylinder at least in part of an operatingrange in which the engine is in the two-cylinder interconnectconfiguration and, when the engine is in the operating range in whichthe two-cylinder interconnect configuration is formed and the engine isstill in a specific low-temperature state in which the temperature inthe following cylinder is judged to have not reached a level suitablefor combustion by compressed self-ignition, the air-fuel ratio in thefollowing cylinder is made substantially equal to the stoichiometricair-fuel ratio to make combustion by forced ignition therein, whereasthe air-fuel ratio in the preceding cylinder is made larger than a casewhere combustion is made by compressed self-ignition in the followingcylinder.

According to this arrangement, a mixture in an entire combustion chamberof the following cylinder burns up in an instant when combustion bycompressed self-ignition is made therein, so that it is possible toprevent delayed combustion which would not produce any work and gain ahigh fuel economy improvement effect. When the engine is still in thelow-temperature state in which the temperature in the following cylinderhas not reached the level suitable for combustion by compressedself-ignition and combustion by forced ignition is made in the followingcylinder, thermal efficiency is improved because the air-fuel ratio inthe preceding cylinder is increased. Since the total amount of fuelinjected into the preceding cylinder and the following cylinder is keptconstant if the amount of intake air remains constant, the amount offuel injected into the following cylinder increases by as much as theamount of fuel injected into the preceding cylinder is reduced forincreasing the air-fuel ratio. Nevertheless, evaporation of the fuel isaccelerated in the following cylinder, resulting in an improvement incombustibility, and pumping loss in the following cylinder becomessmaller than in the preceding cylinder because what is introduced intothe following cylinder is the high-temperature burned gas led from thepreceding cylinder. By increasing the proportion of the fuel combustedin the following cylinder in this manner, it is possible to swiftlyincrease the temperature in the following cylinder and transfer it to astate of combustion by compressed self-ignition soon while achieving afurther improvement in overall fuel economy.

Preferably, the air-fuel ratio in the preceding cylinder is set suchthat the excess-air factor becomes equal to or larger than 3 andstratified charge combustion is made in the preceding cylinder when theengine is in the specific low-temperature state.

According to this arrangement, it is possible to significantly improvethermal efficiency and obtain a remarkable fuel economy improvementeffect by making the excess-air factor in the preceding cylinder equalto or larger than 3 to produce an extremely “lean” air-fuel ratio ofapproximately 50, for example. Compared to a case where one-half of thetotal amount of fuel is supplied to the following cylinder, for example,the amount of fuel supplied to the following cylinder is increased by30% or more in this arrangement. This provides a great contribution tofuel economy improvement and serves to quickly increase the temperaturein the following cylinder. It is to be pointed out that even if thefollowing cylinder is set to such an extremely “lean” air-fuel ratio, itis possible to produce a state of stable combustion in the followingcylinder by making stratified charge combustion in which fuelconcentration around a spark plug at an ignition point is increased.

Preferably, the air-fuel ratio in the preceding cylinder is maderelatively large in a specific low-load region of the operating range inwhich the engine is in the two-cylinder interconnect configurationcompared to a higher-load region.

According to this arrangement, the lower the engine load, the larger theair-fuel ratio in the preceding cylinder is made in the aforementionedspecific low-load region, resulting in a further improvement in fueleconomy.

Preferably, control operation for the aforementioned specificlow-temperature state should be performed when the engine is at or closeto its idling speed. This makes it possible to achieve stable combustionfree of misfire and produce a high fuel economy improvement effect byswiftly increasing the temperature in the following cylinder even whenthe engine is in the low-load, low-speed operating range in which theengine is at or close to its idling speed.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a general plan view of an entire engine provided with acontrol device according to an embodiment the invention;

FIG. 2 is a schematic cross-sectional view of an engine body andassociated elements;

FIG. 3 is a block diagram of a control system;

FIG. 4 is an explanatory diagram showing engine operating ranges;

FIG. 5 is a cross-sectional front view showing a specific constructionof a valve stop mechanism;

FIG. 6 is a cross-sectional plan view showing the specific constructionof the valve stop mechanism;

FIG. 7 is a perspective diagram showing the specific construction of acenter tappet and a side tappet;

FIG. 8 is a cross-sectional front view showing an example of anotherspecific construction of the valve stop mechanism;

FIG. 9 is a diagram showing timing of exhaust strokes and intake strokesas well as fuel injection timing and ignition timing of individualcylinders;

FIG. 10 is an explanatory diagram showing substantial fresh air and gasflow paths formed in a low-load, low-speed operating range;

FIG. 11 is an explanatory diagram showing substantial fresh air and gasflow paths formed in a high-load or high-speed operating range;

FIG. 12 is a diagram showing another example of timing of fuel injectioninto following cylinders;

FIG. 13 is a diagram showing still another example of timing of fuelinjection into the following cylinders;

FIG. 14 is a block diagram of a control system according to anotherembodiment;

FIG. 15 is a time chart showing timing of switching from specialoperation mode to normal operation mode;

FIG. 16 is a time chart showing switching from the normal operation modeto the special operation mode;

FIG. 17 is a flowchart showing operation mode switching controloperation;

FIG. 18 is a block diagram of a control system according to stillanother embodiment;

FIG. 19 is a block diagram showing the functional configuration of afuel controller shown in FIG. 18;

FIG. 20 is a block diagram of a control system according to yet anotherembodiment;

FIG. 21 is an explanatory diagram showing engine operating ranges;

FIG. 22 is a diagram showing timing of exhaust strokes and intakestrokes as well as fuel injection timing and ignition timing of theindividual cylinders;

FIG. 23 is a diagram showing excess-air factors in the individualcylinders;

FIG. 24 is a graph showing the amounts of fuel injected into theindividual cylinders;

FIG. 25 is a graph showing the relationship between engine load andexcess-air factor;

FIG. 26 is a general plan view showing the construction of intake andexhaust ports, intercylinder gas channels and associated elementsaccording to another embodiment;

FIG. 27 is a general plan view showing a flow path switcher andassociated elements according to another embodiment;

FIG. 28 is an explanatory diagram showing switchable periods duringwhich on-off valves may be switched when engine operating condition haschanged according to the embodiment of FIG. 27;

FIG. 29 is a general plan view showing a flow path switcher andassociated elements according to still another embodiment;

FIG. 30 is a general plan view showing an embodiment in which fuelinjectors are provided in intake channels of the following cylinders;and

FIG. 31 is a general plan view showing an embodiment provided with aturbocharger.

BEST MODE FOR CARRYING OUT THE INVENTION

Embodiments of the present invention are now described with reference tothe drawings.

FIG. 1 shows the general construction of an engine provided with acontrol device according to an embodiment the invention, and FIG. 2generally shows the construction of one cylinder of an engine body,intake and exhaust valves provided in the cylinder, etc. Referring tothese Figures, the engine body 1 has a plurality of cylinders.Specifically, it has four cylinders designated 2A to 2D in theillustrated embodiment, with one each piston 3 fitted in the individualcylinders 2A–2D and a combustion chamber 4 formed above the piston 3.

There is provided a spark plug 7 at the top of the combustion chamber 4in each cylinder 2 in such a way that a far end of the spark plug 7 islocated inside the combustion chamber 4. The spark plug 7 is connectedto an ignition circuit 8 which permits electronic control of ignitiontiming.

On one side of the combustion chamber 4 of each cylinder 2, there isprovided a fuel injector 9 for injecting fuel directly into thecombustion chamber 4. The fuel injector 9 incorporates a needle valveand a solenoid which are not illustrated. Driven by a later-describedpulse signal input, the fuel injector 9 causes its needle valve to openat pulse input timing during a period corresponding to the pulselengthof the pulse signal and injects a specific amount of fuel determined bythe valve open period. Although not illustrated, the fuel is suppliedfrom a fuel pump to the fuel injector 9 through a fuel-feeding passage,a fuel-feeding system being so constructed to provide a fuel pressurehigher than the internal pressure of the combustion chamber 4 in eachcompression stroke.

Intake ports 11, 11 a, 11 b and exhaust ports 12, 12 a, 12 b open to thecombustion chambers 4 of the individual cylinders 2A–2D. An intakepassage 15 and an exhaust passage 20 are connected to these ports whichare opened and closed by intake valves 31, 31 a, 31 b and exhaust valves32, 32 a, 32 b, respectively.

The individual cylinders 2A–2D go through successive cycles of intake,compression, expansion and exhaust strokes with specific phase delays.In the four-cylinder engine, of which cylinders 2A–2D are hereinafterreferred to as the first cylinder 2A, the second cylinder 2B, the thirdcylinder 2C and the fourth cylinder 2D as viewed from one end ofcylinder bank, the aforementioned cycles are carried out in the order ofthe first cylinder 2A, the third cylinder 2C, the fourth cylinder 2D andthe second cylinder 2B with a successive phase delay of 180° as shown inFIG. 9, in which “EX” designates exhaust strokes, “IN” designates intakestrokes, “F” designates fuel injection and “S” designates ignition.

There is provided an intercylinder gas channel 22 between two cylindersof which exhaust and intake strokes overlap so that already burned gascan be led from the cylinder in the exhaust stroke (hereinafter referredto as the preceding cylinder in the present Description of theinvention) to the cylinder in the intake stroke (hereinafter referred toas the following cylinder in the present Description of the invention)during a period of overlap of the exhaust and intake strokes. In thefour-cylinder engine of the present embodiment, the exhaust stroke (EX)of the first cylinder 2A overlaps the intake stroke (IN) of the secondcylinder 2B and the exhaust stroke (EX) of the fourth cylinder 2Doverlaps the intake stroke (IN) of the third cylinder 2C as shown inFIG. 9. Therefore, the first cylinder 2A and the second cylinder 2Bconstitute one cylinder pair while the fourth cylinder 2D and the thirdcylinder 2C constitute another cylinder pair, the first cylinder 2A andthe fourth cylinder 2D being the preceding cylinders and the secondcylinder 2B and the third cylinder 2C being the following cylinders.

Specifically, the intake and exhaust ports 11, 11 a, 11 b, 12, 12 a, 12b of the individual cylinders 2A–2D, the intake and exhaust channels 15,20 and the intercylinder gas channels 22 connected to them areconfigured as described in the following.

The intake ports 11 for introducing fresh air, the first exhaust ports12 a for letting out burned gas (exhaust gas) into the exhaust passage20 and the second exhaust ports 12 b for delivering the burned gas tothe respective following cylinders are provided in the first cylinder 2Aand the fourth cylinder 2D which are the preceding cylinders. In thesecond cylinder 2B and the third cylinder 2C which are the followingcylinders, there are provided the first intake ports 11 a forintroducing fresh air, the second intake ports 11 b for introducing theburned gas from the preceding cylinders 2A, 2D and the exhaust ports 12for letting out the burned gas into the exhaust passage 20 are provided.

In the example shown in FIG. 1, two each intake ports 11 are provided inthe first and fourth cylinders 2A, 2D and two each first intake ports 11a are provided in the second and third cylinders 2B, 2C in parallelarrangement on left half sides of the respective combustion chambers 4as illustrated. Also, one each first exhaust port 12 a and secondexhaust port 12 b are provided in the first and fourth cylinders 2A, 2Dand one each second intake port 11 b and exhaust port 12 are provided inthe second and third cylinders 2B, 2C in parallel arrangement on righthalf sides of the respective combustion chambers 4 as illustrated.

Downstream ends of individual intake channels 16 branched out from theintake passage 15 are connected to the intake ports 11 of the first andfourth cylinders 2A, 2D or to the first intake ports 11 a of the secondand third cylinders 2B, 2C. Close to the downstream ends of theindividual branched intake channels 16, there are provided multiplethrottle valves 17 which are interlocked by a common shaft. The multiplethrottle valves 17 are driven by an actuator 18 according to a controlsignal to regulate the amount of intake air. An airflow sensor 19 fordetecting the amount of intake air is provided in the common intakepassage 15 upstream of its joint portion.

Upstream ends of individual exhaust channels 21 branched from theexhaust passage 20 are connected to the first exhaust ports 12 a of thefirst and fourth cylinders 2A, 2D or to the exhaust ports 12 of thesecond and third cylinders 2B, 2C. The intercylinder gas channels 22 areprovided between the first cylinder 2A and the second cylinder 2B andbetween the third cylinder 2C and the fourth cylinder 2D. Upstream endsof the intercylinder gas channels 22 are connected to the second exhaustports 12 b of the first and fourth cylinders 2A, 2D which are thepreceding cylinders while downstream ends of the intercylinder gaschannels 22 are connected to the second intake ports 11 b of the secondand third cylinders 2B, 2C which are the following cylinders.

An O₂ sensor 23 (serving as an exhaust gas concentration detector fordetecting the stoichiometric air-fuel ratio) is provided at a jointportion of the exhaust passage 20, downstream of the individual branchedexhaust channels 21, and a three-way catalyst 24 is provided in theexhaust passage 20 further downstream of the O₂ sensor 23. As iscommonly known, the three-way catalyst 24 is a catalyst which exhibitshigh converting performance with respect to HC, CO and NOx when theair-fuel ratio of the exhaust gas is approximately equal to thestoichiometric air-fuel ratio (i.e., excess-air factor λ=1).

Provided for detecting the air-fuel ratio by measuring the concentrationof oxygen in the exhaust gas, the aforementioned O₂ sensor 23 is formedof a λO₂ sensor whose output suddenly changes particularly at around thestoichiometric air-fuel ratio.

A linear O₂ sensor 25 (serving as an exhaust gas concentration detectorfor detecting a “lean” air-fuel ratio) whose output linearly varies withchanges in oxygen concentration in the exhaust gas is provided in eachof the aforementioned intercylinder gas channels 22.

The intake and exhaust valves for opening and closing the aforementionedintake and exhaust ports of the individual cylinders 2A–2D and a valveactuating mechanism for controlling them are constructed as follows.

The intake ports 11, the first exhaust ports 12 a and the second exhaustports 12 b of the first and fourth cylinders 2A, 2D are provided withthe intake valves 31, the first exhaust valves 32 a and the secondexhaust valves 32 b, respectively. Similarly, the first intake ports 11a, the second intake ports 11 b and the exhaust ports 12 of the secondand third cylinders 2B, 2C are provided with the first intake valves 31a, the second intake valves 31 b and the exhaust valves 32,respectively. The valve actuating mechanism including respectivecamshafts 33, 34 drives these intake and exhaust valves to open andclose them with specific timing such that the exhaust and intake strokesof the individual cylinders 2A–2D are performed with the specific phasedelays mentioned above.

Among the aforementioned intake and exhaust valves, the first exhaustvalves 32 a, the second exhaust valves 32 b and the first intake valves31 a are individually provided with valve stop mechanisms 35 forswitching the respective valves between activated and deactivatedstates. Each of these valve stop mechanisms 35, provided on tappetsfitted between cams and valve stems of the camshafts 33, 34, forexample, can switch the relevant valve between a state in which motionof the cam is transmitted to the valve causing it to open and close inaccordance with supplying and drawing of hydraulic oil and a state inwhich no motion of the cam is transmitted causing it to becomeimmovable. The construction of the valve stop mechanisms 35 will belater described more specifically.

A first control valve 37 is provided in a passage 36 for supplying anddrawing the hydraulic oil to and from the valve stop mechanisms 35 ofthe first exhaust valves 32 a and those of the first intake valves 31 a,while a second control valve 39 is provided in a passage 38 forsupplying and drawing the hydraulic oil to and from the valve stopmechanisms 35 of the second exhaust valves 32 b and those of the secondintake valves 31 b (see FIG. 3).

FIG. 3 shows the construction of a drive/control system. Referring tothis Figure, signals output from the airflow sensor 19, the O₂ sensor 23and the linear O₂ sensors 25 are input to an ECU (electronic controlunit) 40 including a microcomputer for controlling the engine. Alsoinput to the ECU 40 are signals from an engine speed sensor 51 fordetecting engine speed for judging operating condition of the engine andan accelerator pedal stroke sensor 52 for detecting throttle opening(the amount of depression of an accelerator). On the other hand, the ECU40 outputs control signals to the individual fuel injectors 9, theactuator 18 of the multiple throttle valves 17 as well as to the firstand second control valves 37, 39.

The ECU 40 includes an operating condition identifier 41, a valve stopmechanism controller 42, an intake air quantity controller 43 and a fuelinjection controller 44.

The operating condition identifier 41 examines the operating conditionof the engine (engine speed and load) based on the signals from theengine speed sensor 51 and the accelerator pedal stroke sensor 52 andjudges whether the engine operating condition falls in an operatingrange A on a low-load, low-speed side or in an operating range B on ahigh-load, high-speed side shown in FIG. 4.

The valve stop mechanism controller 42 controls the individual valvestop mechanisms 35 as follows by controlling each of the aforementionedcontrol valves 37, 39 depending on whether the engine operatingcondition falls in the operating range A on the low-load, low-speed sideor in the operating range B on the high-load, high-speed side.

-   Operating range A: The first exhaust valves 32 a and the first    intake valves 31 a are in the deactivated state while the second    exhaust valves 32 b and the second intake valves 31 b are in the    activated state.-   Operating range B: The first exhaust valves 32 a and the first    intake valves 31 a are in the activated state while the second    exhaust valves 32 b and the second intake valves 31 b are in the    deactivated state.

The valve stop mechanism controller 42 and the individual valve stopmechanisms 35 controlled by it together constitute a flow path switcherfor switching gas flow paths as will be later described in great detail.

The intake air quantity controller 43 controls the opening of eachthrottle valve 17 (throttle opening) by controlling the actuator 18. Theintake air quantity controller 43 determines a target intake airquantity from a map, for example, based on the engine operatingcondition and controls the throttle opening according to the targetintake air quantity obtained. In the low-load, low-speed operating rangeA, introduction of intake air to the following cylinders (second andthird cylinders 2B, 2C) through the branched intake channels 16 isinterrupted and excess air in the gas introduced from the precedingcylinders is used for combustion as will be later described. Therefore,the throttle opening is regulated in such a manner that air necessaryfor combustion of fuel in a total of two preceding and followingcylinders is supplied to the preceding cylinders (first and fourthcylinders 2A, 2D).

The aforementioned fuel injection controller 44 serves to control theamounts of fuel to be injected from the fuel injectors 9 provided in theindividual cylinders 2A–2D and injection timing according to the engineoperating condition. In particular, the fuel injection controller 44differently controls fuel injection depending on whether the engineoperating condition falls in the operating range A or in the operatingrange B.

Specifically, when the engine operating condition falls in the operatingrange A on the low-load, low-speed side, the fuel injection controller44 controls the amounts of fuel injected into the preceding cylinders(first and fourth cylinders 2A, 2D) such that the air-fuel ratio becomeslarger than the stoichiometric air-fuel ratio, preferably approximatelyequal to twice the stoichiometric air-fuel ratio or larger, to create alean mixture, and sets injection timing to inject the fuel during thecompression stroke to thereby produce stratified charge combustion inthe preceding cylinders 2A, 2D. On the other hand, the fuel injectioncontroller 44 controls the amounts of fuel injected into the followingcylinders (second and third cylinders 2B, 2C) to obtain thestoichiometric air-fuel ratio therein by feeding the fuel into burnedgas of a “lean” air-fuel ratio introduced from the preceding cylinders2A, 2D, and sets injection timing to enable ignition and combustion inan atmosphere rich in burned gas. As an example, the fuel is injectedduring the compression stroke to ensure ignitability.

The aforementioned control operation of the amounts of injected fuel isperformed by feedback control based on the outputs from the airflowsensor 19, the O₂ sensor 23, etc. Specifically, a basic fuel injectionquantity is so calculated for each cylinder based on the amount ofintake air as to produce a specific “lean” air-fuel ratio in thepreceding cylinders 2A, 2D and the stoichiometric air-fuel ratio in thefollowing cylinders 2B, 2C. The amounts of fuel injected into thepreceding cylinders 2A, 2D are corrected by feedback based on theoutputs from the linear O₂ sensors 25 provided in the intercylinder gaschannels 22, and the amounts of fuel injected into the followingcylinders 2B, 2C are corrected by feedback based on the output from theO₂ sensor 23 provided in the exhaust passage 20.

When the engine operating condition falls in the operating range B onthe high-load, high-speed side, on the other hand, the amounts ofinjected fuel are so controlled as to produce an air-fuel ratio equal toor larger than the stoichiometric air-fuel ratio in the individualcylinders 2A–2D. For example, the air-fuel ratio is made equal to thestoichiometric air-fuel ratio in most regions of the operating range Band made lower than the stoichiometric air-fuel ratio to produce a richmixture at and around a full-throttle load region. In this engineoperating condition, the injection timing is so set as to produceuniform charge combustion by injecting the fuel in the intake strokeinto the individual cylinders 2A–2D.

FIGS. 5–7 show a specific example of the mechanism (the valve stopmechanism 35 shown in FIGS. 1–3) for switching the intake and exhaustvalves between the activated and deactivated states, and FIG. 8 shows analternative specific example of the above mechanism.

Referring to these Figures, each camshaft of the valve actuatingmechanism for driving the intake and exhaust valves is provided with afirst cam 33 for opening and closing the intake or exhaust valve bylifting it and a pair of second cams 34 for retaining the intake orexhaust valve in a closed state by preventing it from being lifted.There is provided a mechanism for switching the camshaft between a statein which motion of the first cam 33 is transmitted to the valve and astate in which no motion of the is transmitted between these cams 33, 34and the intake or exhaust valve. A mechanism 35 a shown in FIGS. 5–7includes a center tappet 61 provided at a position corresponding to thefirst cam 33 of the valve actuating mechanism, a side tappet 62 having apair of projecting parts 63 located at positions corresponding to thesecond cams 34, and a pair of compression coil springs 64 locatedbetween the bottom of the side tappet 62 and the bottom of the centertappet 61 for pushing the center tappet 61 in a direction to keep itstop surface in tight contact with the first cam 33.

In the center tappet 61 and the two projecting parts 63 of the sidetappet 62, there are formed lock holes 65, 66 at corresponding positionsso that these lock holes 65, 66 align to form a single through hole whenthe center tappet 61 is in its upper position shown in FIG. 5. A lockpin 67 having a flange 67 a is fitted slidably in its axial direction inthe lock hole 65 of the center tappet 61. A first holder 68 having acavity for accommodating a first plunger 75 and a pusher 76 formed of acompression coil spring for forcing the first plunger 75 against thelock pin 67 is fitted in the lock hole 66 formed in one of theprojecting parts 63 of the side tappet 62, while a second holder 70 forretaining a second plunger 69 is fitted in the lock hole 66 formed inthe other projecting part 63.

First and second bushes 71, 72 for holding both ends of the lock pin 67and a pusher 73 formed of a compression coil spring for forcing the lockpin 67 in the direction of its root end (toward the second plunger 69)are fitted in the lock hole 65 of the center tappet 61. Under normalconditions, the flange 67 a of the lock pin 67 is supported at its homeposition where the flange 67 a is in contact with a far end of thesecond bush 72 due to pushing forces of the pushers 73, 76 as shown inFIGS. 5 and 6, so that the lock pin 67 is accommodated bridging a gapbetween the lock hole 65 of the center tappet 61 and the second holder70, and the first plunger 75 is accommodated bridging the first holder68 and the first bush 71, whereby the center tappet 61 and the sidetappet 62 are held in an interlocked condition. As a result, a drivingforce of the center tappet 61 which is activated by the first cam 33 istransmitted to a stem end 74 of the intake or exhaust valve through theside tappet 62, thereby causing the intake or exhaust valve to open andclose.

When the hydraulic oil is supplied from a later-described passage 36 afor supplying and drawing the hydraulic oil to a line between a root endof the second plunger 69 and the bottom of the second holder 70, thelock pin 67 forced toward the first plunger 75 moves in a directionshown by an arrow in FIG. 5 up to an activated position where the lockpin 67 is accommodated in the lock hole 65 of the center tappet 61overwhelming the pushing force of the pusher 73, and the first plunger75 forced toward the first holder 68 moves in a direction shown by anarrow in FIG. 5 up to an activated position where first plunger 75 isaccommodated in the first holder 68 overwhelming the pushing force ofthe pusher 76, whereby the center tappet 61 is disengaged from the sidetappet 62. As a result, transmission of a driving force of the first cam33 to the intake or exhaust valve through the side tappet 62 isinterrupted, thereby holding the intake or exhaust valve in its closedstate.

A mechanism 35 b shown in FIG. 8 is constructed in a similar fashion asthe mechanism 35 a shown in FIGS. 5–7 except that the center tappet 61is disengaged from the side tappet 62 under normal conditions with thelock pin 67 accommodated in the lock hole 65 of the center tappet 61 andthe second plunger 69 held at its home position where it is accommodatedin the second holder 70.

Under normal conditions, supply of the hydraulic oil to the secondswitching mechanism 35 b is interrupted so that the relevant intake orexhaust valve is held in its closed state. Also, as the hydraulic oil issupplied from the later-described passage 36 a for supplying and drawingthe hydraulic oil to the line between the root end of the second plunger69 and the bottom of the second holder 70, a far end of the secondplunger 69 goes into the lock hole 65 of the center tappet 61, and thelock pin 67 forced by the second plunger 69 is pushed toward the firstholder 68 overwhelming the pushing force of the pusher 73 and moves upto the activated position where a far end of the lock pin 67 is locatedwithin the first holder 68 as shown in FIG. 8, whereby the center tappet61 and the side tappet 62 are joined together. As a result, the drivingforce of the first cam 33 is transmitted to the stem end 74 of theintake or exhaust valve through the side tappet 62, thereby causing theintake or exhaust valve to open and close.

When the aforementioned specific example is applied to the valve stopmechanism 35 of FIGS. 1–3, the mechanism 35 a of FIGS. 5–7 shouldpreferably be adopted to the first exhaust valves 32 a of the precedingcylinders 2A, 2D and the first intake valves 31 a of the followingcylinders 2B, 2C while the mechanism 35 b of FIG. 8 should preferably beadopted to the second exhaust valves 32 b of the preceding cylinders 2A,2D and the second intake valves 31 b of the following cylinders 2B, 2C,for example.

With this arrangement, the aforementioned first exhaust valves 32 a andfirst intake valves 31 a are automatically set in a state in which thedriving force of the first cams 33 is transmitted to the valves 31 a, 32a, and the aforementioned second exhaust valves 32 b and second intakevalves 31 b are automatically set in a state in which the driving forceof the first cams 33 is not transmitted to the valves 31 b, 32 b underengine stop conditions. It is therefore possible to produce a state inwhich fresh air is introduced into the individual cylinders 2A–2Dwithout the need to supply hydraulic oil pressure to the aforementionedmechanisms 35 a, 35 b at engine startup. This serves to ensure ease ofengine startup.

Operational effects of the aforementioned device of the presentembodiment are now described with reference to FIGS. 9–11.

In the operating range A on the low-load, low-speed side, the firstexhaust valves 32 a and the first intake valves 31 a are in thedeactivated state while the second exhaust valves 32 b and the secondintake valves 31 b are in the activated state as stated above, so thatsubstantial fresh air and gas flow paths as shown in FIG. 10 are formed.As a result, there is formed a dual two-cylinder interconnectconfiguration, in which the burned gas discharged from the precedingcylinders (first and fourth cylinders) 2A, 2D is introduced directlyinto the respective following cylinders (second and third cylinders) 2B,2C through the intercylinder gas channels 22 and the burned gasdischarged from only the following cylinders 2B, 2C is led to theexhaust passage 20 associated with the three-way catalyst 24.

In this condition, fresh air is introduced through the intake passage 15(arrows “a” in FIG. 10) into the preceding cylinders 2A, 2D in theirintake stroke, the fuel is injected into the preceding cylinders 2A, 2Din their compression stroke with the amounts of injected fuelfeedback-controlled such that the air-fuel ratio detected by each linearO₂ sensor 25 becomes equal to the aforementioned specific “lean”air-fuel ratio, and the mixture is ignited at specific ignition pointsto produce stratified charge combustion at the “lean” air-fuel ratio(refer to FIG. 9).

Subsequently, the burned gas discharged from the preceding cylinders 2A,2D is introduced into the respective following cylinders 2B, 2C throughthe intercylinder gas channels 22 during periods when the exhaust strokeof the preceding cylinders 2A, 2D overlaps the intake stroke of thefollowing cylinders 2B, 2C (open arrows in FIG. 9 and arrows “b” in FIG.10). The fuel is injected with appropriate timing (during thecompression stroke, for example) into the following cylinders 2B, 2C,with the amounts of injected fuel controlled based on the output of theO₂ sensor 23, to achieve the stoichiometric air-fuel ratio by acombination of the burned gas of the “lean” air-fuel ratio introducedfrom the preceding cylinders 2A, 2D and the newly supplied fuel, and themixture is ignited at specific ignition points to produce combustion inthe following cylinders 2B, 2C (refer to FIG. 9). After combustion inthe following cylinders 2B, 2C, the resultant burned gas is dischargedinto the exhaust passage 20 associated with the three-way catalyst 24(arrows “c” in FIG. 10).

Since stratified charge combustion is performed at the “lean” air-fuelratio in the preceding cylinders 2A, 2D as stated above, thermalefficiency is improved and pumping loss is reduced in the precedingcylinders 2A, 2D, and a combined effect thereof results in a significantimprovement in fuel economy. In the following cylinders 2B, 2C, on theother hand, the mixture is combusted while being controlled to thestoichiometric air-fuel ratio with additional fuel supplied to theburned gas in an excess-air state. Therefore, although the thermalefficiency of the following cylinders 2B, 2C is slightly low compared tothe preceding cylinders 2A, 2D in which the stratified charge combustionis performed at the “lean” air-fuel ratio, it is possible to achieve asufficient effect of improving the fuel economy due to a reduction inpumping loss.

In addition, it is not necessary to provide a lean NOx catalyst unlikeconventional lean burn engines and sufficient emission-cleaningperformance is ensured with the three-way catalyst 24 alone, because thegas discharged from the following cylinders 2B, 2C into the exhaustpassage 20 is at the stoichiometric air-fuel ratio.

As it is not necessary to provide the lean NOx catalyst as stated above,there is no need to temporarily lower the air-fuel ratio foraccelerating release and reduction of NOx when the amount of NOxadsorbed by the lean NOx catalyst has increased, whereby a decrease infuel economy improvement effect can be avoided. Moreover, thearrangement of the embodiment does not cause the earlier-mentionedproblem of poisoning of the lean NOx catalyst by sulfurization.

Furthermore, NOx emission is sufficiently decreased in this embodiment.This is because the air-fuel ratio in the preceding cylinders 2A, 2D ismade approximately equal to twice the stoichiometric air-fuel ratio orlarger by keeping the amount of NOx generated in these cylinders 2A, 2Dto a relatively low level, and the burned gas is introduced from thepreceding cylinders 2A, 2D into the following cylinders 2B, 2C to createa condition equivalent to what would occur when a great deal of exhaustgas is introduced by EGR operation. The arrangement of the embodiment isadvantageous for improving the quality of emissions from such a point ofview as well.

The burned gas is introduced from the preceding cylinders 2A, 2D intothe following cylinders 2B, 2C through the respective intercylinder gaschannels 22 as stated above. It is possible to regulate the amount ofdissipated heat by adjusting the length of each intercylinder gaschannel 22, for example, and thereby regulate the temperature of theburned gas introduced into the following cylinders 2B, 2C. By regulatingthe temperature of the burned gas in this way and also adjusting thetiming of fuel injection into the following cylinders 2B, 2Cappropriately, it is possible to maintain good ignitability andcombustibility in the following cylinders 2B, 2C as well into which agreat deal of exhaust gas is introduced.

Although the combustibility in the following cylinders 2B, 2C lessenswhen the ratio of excess air to the gas introduced from the precedingcylinders 2A, 2D into the following cylinders 2B, 2C decreases, thecombustibility in the following cylinders 2B, 2C is maintained as longas the air-fuel ratio in the preceding cylinders 2A, 2D is madeapproximately equal to twice the stoichiometric air-fuel ratio orlarger.

In the operating range B on the high-load, high-speed side, on the otherhand, the first exhaust valves 32 a and the first intake valves 31 a arein the activated state while the second exhaust valves 32 b and thesecond intake valves 31 b are in the deactivated state as stated above,so that substantial fresh air and gas flow paths as shown in FIG. 11 areformed. As a result, the intake ports 11, 11 a and the exhaust ports 12,12 a of the individual cylinders 2A–2D work substantially independentlyof one another, so that fresh air is introduced through the intakepassage 15 and the intake ports 11, 11 a into the respective cylinders2A–2D and the burned gas is discharged from the cylinders 2A–2D into theexhaust passage 20 through the respective exhaust ports 12, 12 a. Inthis operating range B, the amount of intake air and the amounts ofinjected fuel are so controlled that the air-fuel ratio becomes equal toor smaller than the stoichiometric air-fuel ratio to produce a richmixture to maintain engine output performance.

Although stratified charge combustion is performed at the stoichiometricair-fuel ratio in the following cylinders 2B, 2C by setting a fuelinjection point in the compression stroke in the foregoing embodiment,multi-point fuel injection (F1, F2) may be performed for each of thefollowing cylinders 2B, 2C by injecting the fuel twice during the intakestroke and the compression stroke as depicted in FIG. 12. It would bepossible by doing so to prevent excessive concentration of fuel aroundthe spark plug 7 and produce combustion under slightly stratifiedconditions.

If it is possible to maintain ignitability even when the fuel isuniformly dispersed in the following cylinders 2B, 2C due to asufficiently high temperature of the burned gas introduced from thepreceding cylinders 2A, 2D into the following cylinders 2B, 2C, forexample, one-point fuel injection may be performed for each of thefollowing cylinders 2B, 2C by injecting the fuel only once during theintake stroke as depicted in FIG. 13.

FIG. 14 shows a control system according to another embodiment theinvention. In this embodiment, the engine is switched between anoperation mode in which gas flow paths (refer to FIG. 1) are connectedto form a dual two-cylinder interconnect configuration (refer to FIG.10) and combustion is produced at a “lean” air-fuel ratio in thepreceding cylinders 2A, 2D while combustion is produced at thestoichiometric air-fuel ratio in the following cylinders 2B, 2C (thisoperation mode is hereinafter referred to as the special operation mode)and an operation mode in which combustion is produced at a specificair-fuel ratio with the intake and exhaust ports of the individualcylinders 2A–2D working independently of one another (this operationmode is hereinafter referred to as the normal operation mode) accordingto engine operating conditions, wherein specially devised controloperation is performed at operation mode switching. Referring to FIG.14, an ECU 40′ includes a switching controller 45 in addition to thesame means 41–44 as the ECU 40 shown in FIG. 3. Signals output from anairflow sensor 19, an O₂ sensor 23, linear O₂ sensors 25, an enginespeed sensor 51, an accelerator pedal stroke sensor 52 and a watertemperature sensor 53 are input to the ECU 40′.

The switching controller 45 serves to judge whether the aforementionedgas flow paths have been switched by detecting the period (timeinterval) of intake air pulsations corresponding to the number of intakeair pulsations occurring while a crankshaft of the engine makes aspecific number of turns based on a change in the amount of intake airdetected by an intake air quantity detector formed of the airflow sensor19 and the engine speed detected by the engine speed sensor 51 when theaforementioned operating condition identifier 41 has judged that theoperating range of the engine has changed from the operating range A onthe low-load, low-speed side to the operating range B on the high-load,high-speed side shown in FIG. 4, or vice versa.

In the dual two-cylinder interconnect configuration in which fresh airis introduced into the preceding cylinders 2A, 2D only, two intake airpulsations occur while the crankshaft of the engine makes one turn. Inthe aforementioned independent cylinder configuration in which fresh airis separately introduced into the individual cylinders 2A–2D, on theother hand, four intake air pulsations occur while the crankshaft of theengine makes one turn, so that the period of intake air pulsations issuddenly reduced by approximately half due to this increase in thenumber of intake air pulsations. The switching controller 45 cantherefore judge whether intake air and exhaust gas flow paths have beenswitched by the aforementioned flow path switcher by detecting a changein the period of intake air pulsations. Following a point in time whenthe switching controller 45 has verified that the intake air flow pathshave been switched between the dual two-cylinder interconnectconfiguration and the independent cylinder configuration as a result ofswitching of the intake air and exhaust gas flow paths by the flow pathswitcher, air-fuel ratio control operation corresponding to theoperation mode selected after the flow path switching is performed.

More specifically, after a switching signal is output from the operatingcondition identifier 41 to the flow path switcher at time T1 when theoperating condition identifier 41 judges that the operating range of theengine has transferred from the operating range A on the low-load,low-speed side to the operating range B on the high-load, high-speedside, switching of the flow paths for the aforementioned two cylinders2A, 2B is commenced at time T2 when the intake and exhaust valves of thefirst cylinder 2A which is a preceding cylinder and the intake andexhaust valves of the second cylinder 2B which is a following cylinderare simultaneously set to the closed state as shown in FIG. 15. Then,upon detecting intake air pulsations occurring at time T3 when fresh airis introduced into the second cylinder 2B as a result of execution ofthis switching of the flow paths, the switching controller 45 judgesthat the switching of the intake air flow paths has been completed. Atthis point in time the engine is transferred to a state of air-fuelratio control operation in the normal operation mode in which theair-fuel ratio in the individual cylinders 2A–2D is set to a valueapproximately corresponding to the stoichiometric air-fuel ratio.

The flow paths for the fourth cylinder 2D which is a preceding cylinderand the third cylinder 2C which is a following cylinder other than thosementioned above are switched at time T4 when the intake and exhaustvalves of the two cylinders 2D, 2C are simultaneously set to the closedstate, the time T4 being approximately coinciding with the time T3 whenthe switching controller 45 judges that the switching of the intake airflow paths has been completed. Switching from the special operation modeto the normal operation mode is completed at time T5 when introductionof fresh air into the third cylinder 2C begins upon completion of theaforementioned switching of the flow paths. Preparation for executingair-fuel ratio control operation after the operation mode switching isbegun by gradually increasing the opening of the throttle valves 17 fromthe time T1 when the switching signal is output from the operatingcondition identifier 41 to the flow path switcher.

Also, after a switching signal is output from the operating conditionidentifier 41 to the flow path switcher at time T11 when the operatingcondition identifier 41 judges that the operating range of the enginehas transferred from the operating range B on the high-load, high-speedside to the operating range A on the low-load, low-speed side, switchingof the flow paths for the aforementioned two cylinders 2A, 2B iscommenced at time T12 when the intake and exhaust valves of the firstcylinder 2A which is a preceding cylinder and the intake and exhaustvalves of the second cylinder 2B which is a following cylinder aresimultaneously set to the closed state as shown in FIG. 16. Then, upondetecting a loss of intake air pulsations at time T13 when introductionof fresh air into the second cylinder 2B is interrupted as a result ofexecution of this switching of the flow paths, the switching controller45 judges that the switching of the intake air flow paths has beencompleted. At this point in time the engine is transferred to a state ofair-fuel ratio control operation in the special operation mode in whichthe air-fuel ratio in the individual cylinders 2A–2D is set to a valueapproximately corresponding to the stoichiometric air-fuel ratio.

The flow paths for the fourth cylinder 2D which is a preceding cylinderand the third cylinder 2C which is a following cylinder other than thosementioned above are switched at time T14 when the intake and exhaustvalves of the two cylinders 2D, 2C are simultaneously set to the closedstate, the time T14 being approximately coinciding with the time T13when the switching controller 45 judges that the switching of the intakeair flow paths has been completed.

Engine control operation performed by a control device according to thepresent embodiment is now described with reference to FIG. 17. Afterthis control operation has started, the control device judges whetherthe switching signal has been output from the operating conditionidentifier 41 to the valve stop mechanism controller 42 of the flow pathswitcher following a change in engine operating conditions (step S1). Ata point in time when the judgment result in step S1 proves to be in theaffirmative, preparation for executing air-fuel ratio control operationcorresponding to the operation mode selected after its switching isbegun (step S2). Then, at the timing of the aforementioned flow pathswitching, that is, when the intake and exhaust valves of the precedingand following cylinders 2D, 2C are simultaneously closed, controloperation for switching the intake air and exhaust gas flow paths isexecuted (step S3).

Subsequently, a judgment is made by the switching controller 45 todetermine whether the period of intake air pulsations has suddenlychanged corresponding to a sensing signal of the airflow sensor 19 (stepS4), and the aforementioned air-fuel ratio control operationcorresponding to the operation mode selected after its switching isexecuted at a point in time when the judgment result in step S1 provesto be in the affirmative (step S5). When the engine is switched from thespecial operation mode to the normal operation mode, for example, theamount of intake air and the amounts of injected fuel are controlled insuch a manner that the air-fuel ratio in each of the cylinders 2A–2Dbecomes approximately equal to a value corresponding to thestoichiometric air-fuel ratio. When the engine is switched from thenormal operation mode to the special operation mode, on the other hand,the amount of intake air and the amounts of injected fuel are controlledin such a manner that combustion in the preceding cylinders 2A, 2D isproduced in a lean mixture state in which the air-fuel ratio is largerthan the stoichiometric air-fuel ratio by a specific amount by injectingthe fuel into these cylinders 2A, 2D and combustion in the followingcylinders 2B, 2C is produced with the air-fuel ratio made approximatelyequal to the value corresponding to the stoichiometric air-fuel ratio bysupplying the burned gas drawn from the preceding cylinders 2A, 2D andthe fuel into the following cylinders 2B, 2C.

The above-described device of the present embodiment is so constructedas to judge whether the flow paths have been switched by the flow pathswitcher with reference to a sensing signal output from an intake airpulsation detector formed of the airflow sensor 19 at the aforementionedoperation mode switching and execute the air-fuel ratio controloperation corresponding to the operation mode selected after the flowpath switching following the point in time when the switching of theflow paths has been detected. Thus, when the switching between thespecial operation mode and the normal operation mode is made, theair-fuel ratio control operation corresponding to the operation modeselected after the flow path switching is properly performed aftercompletion of the switching of the flow paths has been verified by theflow path switcher with reference to the aforementioned sensing signal.

At the switching from the special operation mode to the normal operationmode, there could arise such a problem as misfire due to an insufficientamount of fresh air introduced into the following cylinders 2B, 2C, forexample, if the air-fuel ratio control operation for the normaloperation mode, in which combustion is produced with the air-fuel ratioin the individual cylinders 2A–2D made approximately equal to thestoichiometric air-fuel ratio, is performed although the aforementionedindependent cylinder configuration has not been completed yet due to adelay in the flow path switching for some reason. In this embodiment, itis possible to effectively prevent this kind of problem and properlyperform combustion control operation for the normal operation mode uponverifying that the flow paths have been switched to the independentcylinder configuration.

At the switching from the normal operation mode to the special operationmode, there could arise a problem that NOx generated when the air-fuelratios in the preceding cylinders 2A, 2D and the following cylinders 2B,2C increase is led to the exhaust passage 20, for example, if theair-fuel ratio control operation for the special operation mode, inwhich combustion in the preceding cylinders 2A, 2D is produced with theair-fuel ratio made larger than the stoichiometric air-fuel ratio by aspecific amount to create a lean mixture while combustion in thefollowing cylinders 2B, 2C is produced with the air-fuel ratio madeapproximately equal to the stoichiometric air-fuel ratio, is performedalthough the aforementioned dual two-cylinder interconnect configurationhas not been completed yet due to a delay in the flow path switching. Inthis embodiment, it is possible to effectively prevent this kind ofproblem and properly perform combustion control operation for thespecial operation mode upon verifying that the flow paths have beenswitched to the dual two-cylinder interconnect configuration.

Since the device of the present embodiment is so constructed as to judgethat the flow path switching has been completed at a point in time whena sudden change in the period of intake air pulsations is detected withreference to the sensing signal output from the intake air pulsationdetector, it is possible to exactly judge whether or not the flow pathshave been actually switched with reference to the sensing signal outputfrom the existing intake air pulsation detector (airflow sensor 19)provided in the intake passage 15. Moreover, the intake air quantitydetector, that is, the intake air pulsation detector, has an advantagethat it can quickly detect the period of intake air pulsations becauseit detects changes in intake air pulsations transmitted at the soundvelocity.

Furthermore, if the device of the present embodiment is so constructedas to change the intake air and exhaust gas flow paths by varying theamounts of valve lift determined by the valve actuating mechanism fordriving the intake and exhaust valves provided in the individualcylinders 2A–2D by means of the valve stop mechanism controller 42 asstated above, it is possible to quickly switch those flow paths whenswitching the engine between the special operation mode and the normaloperation mode and to properly perform the air-fuel ratio controloperation corresponding to the operation mode selected after the flowpath switching upon verifying that the flow path switching has beencompleted with reference to the sensing signal output from theaforementioned intake air pulsation detector.

Provided with multiple pairs of the preceding and following cylinders ofwhich intake and exhaust strokes overlap with each other, the engine ofthe present embodiment is so constructed as to perform the air-fuelratio control operation corresponding to the operation mode selectedafter the flow path switching in all the pairs of the preceding andfollowing cylinders following the point in time when the switching ofthe flow paths has been first verified in one of the multiple pairs ofthe preceding and following cylinders. Although the flow paths of themultiple pairs of the preceding and following cylinders are switched ina specific order, the air-fuel ratio control operation corresponding tothe operation mode selected after the flow path switching is performedin all the pairs of the preceding and following cylinders at the pointin time when the switching of the flow paths has been verified in thefirst pair of the preceding and following cylinders with reference tothe sensing signal from the aforementioned intake air pulsationdetector. It is therefore possible to quickly and properly execute theair-fuel ratio control operation corresponding to the operation mode.

Furthermore, the device of the present embodiment is so constructed asto judge that the flow paths have been switched at a point in time whenthe switching signal has been output to the flow path switcher followinga change in engine operating conditions and the occurrence of a changein intake air pulsations has been verified with reference to the sensingsignal output from the intake air pulsation detector. This constructionmakes it possible to prevent an incorrect judgment due to sensing errorsof the intake air pulsation detector or noise contained in its sensingsignal, exactly judge that the flow paths have been switched based onthe aforementioned switching signal, and properly perform the air-fuelratio control operation according to the judgment result.

Furthermore, the device of the present embodiment is so constructed asto begin preparation for executing the air-fuel ratio control operationafter the operation mode switching, such as operation for regulating theamount of intake air by varying the opening of each throttle valve 17,for example, at a point in time when it is verified that the flow pathswitching signal has been output to the flow path switcher. Thisconstruction is advantageous in that the air-fuel ratio controloperation corresponding to the operation mode selected after the flowpath switching can be quickly executed upon verifying that the switchingsignal has been output to the flow path switcher following a change inengine operating conditions and the flow paths have been switched as aresult of the occurrence of a change in intake air pulsations accordingto the sensing signal output from the intake air pulsation detector.

FIG. 18 shows a control system according to still another embodiment theinvention. In this embodiment, an ECU 80 includes an operating conditionidentifier 81, a temperature status identifier 82, a mode setter 83, avalve stop mechanism controller 84, an intake air quantity controller85, a fuel controller 86 and an ignition controller 87.

Like the operating condition identifier 41 of FIG. 3, the operatingcondition identifier 81 judges whether the engine operating condition(engine speed and load) falls in the operating range A or B shown inFIG. 4.

The temperature status identifier 82 examines the status of enginetemperature based on a signal output from a water temperature sensor 53.Specifically, it judges whether the engine is in a low temperature rangein which water temperature (engine temperature) is equal to or lowerthan a specific value or in a high temperature range in which the watertemperature is higher than the specific value.

The mode setter 83 selects the special operation mode in theaforementioned operating range A to combust the burned gas introduceddirectly from the preceding cylinders 2A, 2D which are in the exhauststroke into the following cylinders 2B, 2C and the normal operation modein the aforementioned operating range B to combust the mixture in theindividual cylinders 2A–2D independently of one another.

Like the valve stop mechanism controller 42 of FIG. 3, the valve stopmechanism controller 84 controls the individual valve stop mechanisms 35by controlling the aforementioned control valves 37, 39 depending onwhether the engine is in the operating range A or B.

Like the intake air quantity controller 43 of FIG. 3, the intake airquantity controller 85 determines a target intake air quantity from amap, for example, based on the engine operating condition the controlsthe throttle opening according to the target intake air quantityobtained.

The fuel controller 86 performs basically the same control operation asthe fuel injection controller 44 of FIG. 3. Specifically, when thespecial operation mode has been selected, the fuel controller 86controls the amounts of fuel injected into the preceding cylinders 2A,2D such that the air-fuel ratio becomes larger than the stoichiometricair-fuel ratio, preferably approximately equal to twice thestoichiometric air-fuel ratio or larger, to create a lean mixture, andsets injection timing to inject the fuel during the compression stroketo thereby produce stratified charge combustion in the precedingcylinders 2A, 2D. On the other hand, the fuel controller 86 controls theamounts of fuel injected into the following cylinders 2B, 2C to obtainthe stoichiometric air-fuel ratio therein by feeding the fuel intoburned gas of a “lean” air-fuel ratio introduced from the precedingcylinders 2A, 2D, and sets injection timing to enable ignition andcombustion in an atmosphere rich in burned gas. When the normaloperation mode has been selected, the fuel controller 86 controls theamounts of injected fuel to make the air-fuel ratio in the individualcylinders 2A–2D equal to or smaller than the stoichiometric air-fuelratio and sets injection timing to inject the fuel during the intakestroke to thereby produce a uniform mixture.

When the special operation mode is selected, the ratio of the amounts offuel injected into the following cylinders 2B, 2C to the amounts of fuelinjected into the preceding cylinders 2A, 2D is adjusted such that thetotal amount of fuel injected into the two cylinder pairs with respectto the amount of air introduced into the preceding cylinders 2A, 2Dproduces the stoichiometric air-fuel ratio and a balance is achievedbetween a torque generated by the preceding cylinders 2A, 2D and atorque generated by the following cylinders 2B, 2C.

In this respect, the construction of the fuel controller 86 is explainedin greater detail referring to FIG. 19. As shown in FIG. 19, the fuelcontroller 86 includes in its functional configuration a total fuelinjection quantity calculator 86 a, a final fuel injection quantitycalculator 86 b, a torque balance air-fuel ratio setter 86 c, acombustibility judgment section 86 d and a distribution ratio setter 86e.

The total fuel injection quantity calculator 86 a calculates the amountof fuel injected from the fuel injectors 9 based on the amount of intakeair detected by the airflow sensor 19. In particular, when the specialoperation mode has been selected, the total fuel injection quantitycalculator 86 a calculates the sum of the amount of fuel injected intothe preceding cylinders 2A, 2D and the amount of fuel injected into thefollowing cylinders 2B, 2C, or the total fuel injection quantity. Inthis case, the total fuel injection quantity is calculated such that thetotal amount of injected fuel with respect to the amount of airintroduced into the preceding cylinders 2A, 2D produces thestoichiometric air-fuel ratio as stated above.

The final fuel injection quantity calculator 86 b determines a finalquantity of fuel to be injected. When the normal operation mode has beenselected, the final fuel injection quantity calculator 86 b adopts thetotal fuel injection quantity calculated by the total fuel injectionquantity calculator 86 a as the final fuel injection quantity. When thespecial operation mode has been selected, on the other hand, the finalfuel injection quantity calculator 86 b calculates the amounts of fuelinjected into the preceding cylinders 2A, 2D and the following cylinders2B, 2C from the total fuel injection quantity and a later-describeddistribution ratio and adopts these amounts as the final fuel injectionquantity.

The torque balance air-fuel ratio setter 86 c, the combustibilityjudgment section 86 d and the distribution ratio setter 86 e performtheir functions when the special operation mode has been selected.

The torque balance air-fuel ratio setter 86 c calculates the air-fuelratio for the preceding cylinders 2A, 2D from a preprogrammed mapaccording to engine operating conditions (engine speed and load)determined from signals fed from an engine speed sensor 51 and anaccelerator pedal stroke sensor 52. The map is produced byexperimentally predetermining a difference between the torques generatedby the preceding cylinders 2A, 2D and the following cylinders 2B, 2Coccurring due to a difference in thermal efficiency including pumpingloss, for example, and correcting the air-fuel ratio determined bydesign in such a manner that this difference in torque becomes “0”,wherein the air-fuel ratio so determined is a value equal to the “lean”air-fuel ratio larger than the stoichiometric air-fuel ratio, preferablyapproximately equal to twice the stoichiometric air-fuel ratio orlarger, that is calculated from design. The map correlates such valuesto the engine operating conditions.

The combustibility judgment section 86 d judges in advance whether ornot combustion can be normally made based on the ratio of the air-fuelratios calculated by the torque balance air-fuel ratio setter 86 c inaccordance with the engine temperature judged by the signal output fromthe water temperature sensor 53 and the map, and outputs the judgmentresult to the distribution ratio setter 86 e.

The distribution ratio setter 86 e determines the distribution ratio offuel (the total fuel injection quantity mentioned above) to be injectedinto the preceding and following cylinders 2A–2D from the ratio betweenthe air-fuel ratio of the preceding cylinders 2A, 2D and the air-fuelratio (stoichiometric air-fuel ratio) of the following cylinders 2B, 2Cobtained from the aforementioned map. If the judgment result of thecombustibility judgment section 86 d is in the affirmative, that is, ifit judges that normal combustion can be performed, the distributionratio setter 86 e determines the distribution ratio of fuel based on theratio of the air-fuel ratios calculated by the torque balance air-fuelratio setter 86 c and outputs the judgment result to the final fuelinjection quantity calculator 86 b. If the judgment result of thecombustibility judgment section 86 d is in the negative, on the otherhand, that is, if it judges that there is a possibility of misfire orknocking, for instance, the distribution ratio setter 86 e outputs adistribution ratio preset within a range in which normal combustion canbe performed in the preceding and following cylinders 2A–2D, such as adistribution ratio obtained when the air-fuel ratio of the precedingcylinders 2A, 2D is made equal to the value determined by design (i.e.,the value before correction according to the difference in torque), tothe final fuel injection quantity calculator 86 b. In this embodiment,the distribution ratio setter 86 e and the final fuel injection quantitycalculator 86 b together constitute a final fuel injection quantitycontroller.

Operational effects of the aforementioned device of the presentembodiment are now described.

The special operation mode in which combustion is performed in the dualtwo-cylinder interconnect configuration is selected in the operatingrange A on the low-load, low-speed side, while the normal operation modein which combustion is performed with the intake ports 11, 11 a and theexhaust ports 12, 12 a of the individual cylinders 2A–2D workingindependently of one another is selected in the operating range B on thehigh-load, high-speed side in this embodiment as well. Particularly inthe special operation mode, the amounts of fuel injected into thepreceding cylinders 2A, 2D and the following cylinders 2B, 2C arecontrolled such that a balance is achieved between the torques generatedby the preceding and following cylinders 2A–2D. This serves to enhanceNoise, Vibration and Harshness (NVH) performance of the engine in adesirable fashion. Since there exists a difference in thermal efficiencyincluding pumping loss between the preceding cylinders 2A, 2D and thefollowing cylinders 2B, 2C, noise and vibrations are supposed to occurdue to a difference in torque between the cylinders if the amounts offuel injected into the preceding cylinders 2A, 2D and the followingcylinders 2B, 2C are controlled to an equal level. In this embodiment,however, the amounts of injected fuel are controlled as stated above bythe fuel controller 86, so that almost no torque difference occursbetween the preceding cylinders 2A, 2D and the following cylinders 2B,2C and, therefore, it is possible to effectively prevent the occurrenceof vibrations due to the torque difference.

Furthermore, because the fuel controller 86 judges in advance in aprocess of setting the amounts of fuel to be injected (distributionratio) whether combustion can be normally made in the individualpreceding and following cylinders 2A–2D (judgment made by thecombustibility judgment section 86 d) and determines the amounts of fuelto be injected within a range in which normal combustion is possible ifthere is a possibility that normal combustion can not be performed, itis possible to effectively prevent problems, such as the occurrence ofmisfire or knocking, caused by controlling the amounts of fuel to beinjected such that a balance is achieved between the torques generatedby the preceding and following cylinders 2A–2D. More particularly, ifpriority is given to the achievement of a balance of the generatedtorques, the amounts of fuel to be injected (distribution ratio) mightbe set in a range in which the air-fuel ratio in the preceding orfollowing cylinders exceeds a range in which normal combustion ispossible, potentially causing misfire or knocking. In the device of thepresent embodiment, however, the amounts of fuel to be injected aredetermined after the aforementioned judgment has been made in advance,so that the amounts of fuel to be injected into the preceding andfollowing cylinders 2A–2D are continuously controlled to fall within therange in which normal combustion is possible. It is therefore possibleto prevent misfire and knocking and maintain normal operatingconditions.

In one variation of the present embodiment, it is possible to controlthe amounts of fuel to be injected to achieve a balance between thetorques generated by the preceding cylinders 2A, 2D and the followingcylinders 2B, 2C while producing combustion in the following cylinders2B, 2C by compressed self-ignition, using a phenomenon that thetemperature in the following cylinders 2B, 2C rises as a result ofintroduction of the burned gas from the preceding cylinders 2A, 2D.

For example, the fuel controller 86 selects, based on the judgment ofthe engine temperature status, forced ignition mode in which combustionin the following cylinders 2B, 2C is made by forced ignition at lowtemperatures and compressed self-ignition mode in which combustion inthe following cylinders 2B, 2C is made by compressed self-ignition athigh temperatures. Particularly in the compressed self-ignition mode,the fuel controller 86 controls the engine to inject the fuel during theintake stroke. As an alternative, the fuel controller 86 selects thecompressed self-ignition mode in a high-load region of the operatingrange A in which the special operation mode is selected as in alater-described embodiment shown in FIGS. 20–25.

Since the amounts of fuel to be injected into the preceding cylinders2A, 2D and the following cylinders 2B, 2C are so controlled that abalance is achieved between the torques generated by the preceding andfollowing cylinders 2A–2D in the aforementioned compressed self-ignitionmode as well, the NVH performance is enhanced, making it possible toensure desirable operating conditions. In this variation of theembodiment, it is necessary to configure the device such that the torquebalance air-fuel ratio setter 86 c separately stores a map fordetermining the ratio of the air-fuel ratios for the compressedself-ignition mode besides the map for determining the ratio of theair-fuel ratios for the forced ignition mode and the ratio of theair-fuel ratios is set based on the map corresponding to the selectedmode. As will be later described in detail, the thermal efficiency isimproved when compressed self-ignition is performed compared to a casewhere forced ignition is performed. Therefore, the difference betweenthe torques generated by the preceding cylinders 2A, 2D and thefollowing cylinders 2B, 2C is supposed to become even larger in thecompressed self-ignition mode than in the forced ignition mode. Theaforementioned separate maps for the compressed self-ignition mode andthe forced ignition mode are required because it is difficult tocompletely eliminate the difference between the torques generated by thepreceding and following cylinders 2A–2D by just determining the ratio ofthe air-fuel ratios based on the same map as used in the case of forcedignition.

In the foregoing embodiment, values held in the map (i.e., the mapstored in the torque balance air-fuel ratio setter 86 c for setting theratio of the air-fuel ratios) used as the basis for determining theamounts of fuel to be injected into the preceding and followingcylinders 2A–2D are obtained by experimentally determining thedifference between the torques generated by the preceding cylinders 2A,2D and the following cylinders 2B, 2C and by correcting the air-fuelratio determined by design in such a manner that this difference intorque becomes “0”, wherein the air-fuel ratio so determined is a valueequal to the “lean” air-fuel ratio larger than the stoichiometricair-fuel ratio, preferably approximately equal to twice thestoichiometric air-fuel ratio or larger, that is calculated from design.Needless to say, it is possible to use values theoretically obtained (orcalculated) from parameters concerning pumping loss and thermalefficiency of the preceding and following cylinders 2A–2D using designvalues.

FIG. 20 shows a control system according to yet another embodiment theinvention. An ECU 90 shown in this Figure constitutes a controller forproducing combustion with gas flow paths connected to form a dualtwo-cylinder interconnect configuration (refer to FIG. 10) at least in alow-speed, low-load range. It includes an operating condition identifier91, a valve stop mechanism controller 92, an intake air quantitycontroller 93 and a combustion controller 94.

The operating condition identifier 91 examines the operating conditionof the engine (engine speed and load) based on signals fed from anengine speed sensor 51 and an accelerator pedal stroke sensor 52 andjudges whether the engine operating condition falls in an operatingrange A (including subranges A1 and A2) on a low-load, low-speed side orin an operating range B on a high-load, high-speed side shown in FIG.21. Of the operating range A, the operating subrange A1 is a region onthe low-load side, in which the engine is still in a low-temperaturestate where the temperature in the following cylinders 2B, 2C has notreached a level suitable for combustion by compressed self-ignition. Theoperating subrange A1 is made variable with the temperature in thefollowing cylinders 2B, 2C. When the temperature of engine water isrelatively low, for example, the temperature in the following cylinders2B, 2C is also low and, therefore, the operating subrange A1 isexpanded. When the engine water temperature is high, on the contrary,the operating subrange A1 is reduced. The operating range A alsoincludes engine idling conditions. On the other hand, the operatingsubrange A2 is a region of higher load than the operating subrange A1,in which combustion is made by compressed self-ignition in the followingcylinders 2B, 2C.

In principle, the operating condition identifier 91 selects the specialoperation mode in the operating range A to produce combustion in theaforementioned dual two-cylinder interconnect configuration, the normaloperation mode in the operating range B to produce combustion in theaforementioned independent cylinder configuration.

The valve stop mechanism controller 92 and the intake air quantitycontroller 93 are identical in their working to the valve stop mechanismcontroller 42 and the intake air quantity controller 43 shown in FIG. 3,respectively.

The combustion controller 94 Including a fuel injection controller 95and an ignition controller 96, the combustion controller 94 differentlycontrols combustion (i.e., fuel injection and ignition) dependingparticularly on whether the engine is operated in the special operationmode or the normal operation mode.

When the special operation mode is selected, the amounts of fuelinjected into the preceding cylinders (first and fourth cylinders 2A,2D) are so controlled as to produce a “lean” air-fuel ratio larger thanthe stoichiometric air-fuel ratio, the injection timing is so set as toinject the fuel in the compression stroke to produce a stratifiedmixture, and the ignition timing is so set as to perform forced ignitionnear the top dead center in the compression stroke. On the other hand,the amounts of fuel injected into the following cylinders (first andfourth cylinders 2B, 2C) are so controlled as to produce a substantialstoichiometric air-fuel ratio by supplying additional fuel to the burnedgas introduced from the preceding cylinders 2A, 2D, and the injectiontiming is so set as to inject the fuel in the intake stroke. Thecombustion controller 94 causes combustion by forced ignition when theengine is operated in the operating subrange A1 of FIG. 21 and causescombustion by compressed self-ignition when the engine is operated inthe operating subrange A2 of FIG. 21.

FIG. 23 is a graph showing the relationship between the amounts of fuelinjected into the preceding cylinders 2A, 2D and the following cylinders2B, 2C in the special operation mode on condition that the amount ofintake air is unchanged. In FIG. 23, the horizontal axis indicates theamount of fuel F1 injected into each of the preceding cylinders 2A, 2Dand vertical axis indicates the amount of fuel F2 injected into each ofthe following cylinders 2B, 2C. If the amount of fuel to be suppliedinto each of the preceding cylinders 2A, 2D for producing thestoichiometric air-fuel ratio in relation to the amount of intake airintroduced into the same cylinders 2A, 2D is F0, the amounts of fuel F1,F2 have a relationship expressed by the equation F1+F2=F0 as shown inFIG. 23. Thus, the amount of fuel F2 injected into each of the followingcylinders 2B, 2C varies inversely when the amount of fuel F1 injectedinto each of the preceding cylinders 2A, 2D is increased or decreased.

In the higher load region (operating subrange A2 of FIG. 21), a settingat which the amount of fuel F1 injected into each of the precedingcylinders 2A, 2D becomes equal to one-half of the amount of fuel F0 tobe supplied for producing the stoichiometric air-fuel ratio (point Gshown in FIG. 23) is used as a criterion, and the ratio of F1 to F0 isvaried with load. Specifically, the ratio F1/F0 is decreased as the loaddecreases. In the lower load region (operating subrange A1 of FIG. 21),on the other hand, the amount of fuel F1 injected into each of thepreceding cylinders 2A, 2D is set to a value equal to or smaller thanone-third of the amount of fuel F0 to be supplied for producing thestoichiometric air-fuel ratio (point H shown in FIG. 23). In this case,the amount of fuel F2 injected into each of the following cylinders 2B,2C is equal to or larger than two-thirds of the amount of fuel F0.

FIG. 24 is a graph showing the relationship between the amounts of fuelinjected into the preceding cylinders 2A, 2D and excess-air factors inthe preceding and following cylinders 2A–2D. In FIG. 24, the horizontalaxis indicates the amount of fuel F1 injected into each of the precedingcylinders 2A, 2D and vertical axis indicates the excess-air factors inthe individual cylinders 2A–2D. The excess-air factor λ is a parameterindicating how many times as large as the stoichiometric air-fuel ratio(λ=1) is the air-fuel ratio. When there is the relationship F1+F2=F0shown in FIG. 23 between the amounts of fuel injected into theindividual cylinders 2A–2D, the excess-air factor λ in the precedingcylinders 2A, 2D is given by the equation λ=F0/F1 as shown in FIG. 24.As can be seen this relationship, the excess-air factor λ decreases whenthe amount of fuel F1 injected into each of the preceding cylinders 2A,2D is increased, and the excess-air factor λ increases when the amountof fuel F1 injected into each of the preceding cylinders 2A, 2D isdecreased.

When the amount of fuel F1 injected into each of the preceding cylinders2A, 2D is one-half of the amount of fuel F0 to be supplied for producingthe stoichiometric air-fuel ratio (point G shown in FIG. 23), theexcess-air factor λ in the preceding cylinders 2A, 2D becomes 2 (pointG1 shown in FIG. 24). Also, when the amount of fuel F1 injected intoeach of the preceding cylinders 2A, 2D is one-third of the amount offuel F0 to be supplied for producing the stoichiometric air-fuel ratio(point H shown in FIG. 23), the excess-air factor λ in the precedingcylinders 2A, 2D becomes 3 (point H1 shown in FIG. 24). In the operatingsubrange A1 on the low-load side, the excess-air factor is equal to orlarger than 3 (λ≧3). For example, the excess-air factor is set to λ=3.4(air-fuel ratio ≈50).

On the other hand, the amount of fuel F2 injected into each of thefollowing cylinders 2B, 2C is varied inversely when the amount of fuelF1 injected into each of the preceding cylinders 2A, 2D is increased ordecreased as shown in FIG. 23. Therefore, the air-fuel ratio becomessubstantially equal to the stoichiometric air-fuel ratio and theexcess-air factor λ is kept constant at λ=1 in the following cylinders2B, 2C as shown by points G2 and H2 in FIG. 24.

FIG. 25 is a graph showing the relationship between engine load and theexcess-air factor λ in the preceding cylinders 2A, 2D in the specialoperation mode, in which a solid line “a” and a broken line “b” indicateengine characteristics at normal temperatures and low temperatures,respectively. As shown in this Figure, the excess-air factor λ is so setas to become progressively larger (to produce a leaner mixture) as theengine load becomes smaller. Bending points in the graph where theengine characteristics suddenly change correspond to boundaries betweenregions in which forced ignition is executed or compressed self-ignitionis executed in the following cylinders 2B, 2C. Specifically, theexcess-air factor λ in the preceding cylinders 2A, 2D is set to becomelarge particularly when combustion is made by forced ignition in thefollowing cylinders 2B, 2C.

Operational effects of the aforementioned device of the presentembodiment are now described.

The special operation mode is selected in the operating range A on thelow-load, low-speed side, while the normal operation mode is selected inthe operating range B on the high-load, high-speed side in thisembodiment as well. In the special operation mode, the engine is set tothe dual two-cylinder interconnect configuration (refer to FIG. 10). Inthis condition, fresh air is introduced through the intake passage 15into the individual preceding cylinders 2A, 2D in the intake stroke andthe fuel is injected in the compression stroke with the amounts ofinjected fuel feedback-controlled such that the air-fuel ratio becomeslarger than the stoichiometric air-fuel ratio to produce a lean mixturein the preceding cylinders 2A, 2D. Then, the mixture is ignited atspecific ignition points to cause combustion. In the following cylinders2B, 2C, on the other hand, the fuel is injected in the intake stroke,with the amounts of injected fuel controlled such that the air-fuelratio becomes substantially equal to the stoichiometric air-fuel ratioby a combination of the burned gas of a “lean” air-fuel ratio introducedfrom the preceding cylinders 2A, 2D and the newly supplied fuel. Then,combustion is made by forced ignition when the engine is in the lowerload region (operating subrange A1 of FIG. 21), while combustion is madeby compressed self-ignition as a result of pressure and temperaturerises in the combustion chambers 4 near the top dead center in thecompression stroke when the engine is in the higher load region(operating subrange A2 of FIG. 21).

Both thermal efficiency and fuel economy are improved particularly inthe present embodiment. This is because the amount of fuel F1 injectedinto each of the preceding cylinders 2A, 2D is set to a value equal toor smaller than one-third of the total amount of fuel F0 and stratifiedcharge combustion is made at an extremely “lean” air-fuel ratio whichmakes the excess-air factor equal to or larger than 3 (λ≧3). On theother hand, the amount of fuel F2 injected into each of the followingcylinders 2B, 2C is set to a value equal to or larger than two-thirds ofthe total amount of fuel F0 and combustion is made by forced ignitionsubstantially at the stoichiometric air-fuel ratio (λ=1). Since intakeair introduced into the following cylinders 2B, 2C is thehigh-temperature burned gas circulated from the preceding cylinders 2A,2D, evaporation of the fuel is accelerated resulting in an improvementin combustibility and pumping loss in the following cylinders 2B, 2C issmaller than in the preceding cylinders 2A, 2D. As the proportion of thefuel combusted in the following cylinders 2B, 2C so controlled isincreased, the fuel economy is further improved as a whole.

Furthermore, the temperature in the following cylinders 2B, 2C risesrelatively quickly as the amounts of fuel supplied to these cylindersare increased. When the temperature in the following cylinders 2B, 2Chas increased, the operating subrange A1 becomes smaller, so that itbecomes easier for these cylinders to transfer to the operating subrangeA2. For this reason, the following cylinders 2B, 2C can be transferredto a state of combustion by compressed self-ignition so early that astill further improvement in fuel economy can be achieved.

Since the aforementioned control operation is performed close to idlingengine speed as well, it is possible to achieve stable combustion freeof misfire and produce a high fuel economy improvement effect by swiftlyincreasing the temperature in the following cylinders 2B, 2C.

When the special operation mode is selected in the operating subrangeA2, a setting at which the amount of fuel F1 injected into each of thepreceding cylinders 2A, 2D becomes equal to the amount of fuel F2injected into each of the following cylinders 2B, 2C is used as acriterion, and the proportion of the amount of fuel F1 injected intoeach of the preceding cylinders 2A, 2D is increased as the engine loadincreases. For this reason, the temperature of the burned gas introducedinto the following cylinders 2B, 2C is apt to further increase, and thisresults in an increase in ignitability by compressed self-ignition. Onthe other hand, the temperature in the following cylinders 2B, 2C hasalready been sufficiently increased, combustion by compressedself-ignition is performed. Thus, the mixture in the entire combustionchambers 4 of the following cylinders 2B, 2C burns up in an instant. Itis therefore possible to prevent delayed combustion which would notproduce any work and gain a high fuel economy improvement effect.

In addition, since the air-fuel ratio in the preceding cylinders 2A, 2Dis set to progressively larger values (to produce a leaner mixture) asthe engine load decreases in the special operation mode, it is possibleto obtain the fuel economy improvement effect without reversing thetendency of change of the air-fuel ratio with respect to changes in theengine load. This facilitates the control operation and helps achievestable combustion.

Although the operating subrange A1 shown in FIG. 21 is judged to be thelow-temperature state in which the temperature in the followingcylinders 2B, 2C has not reached the level suitable for combustion bycompressed self-ignition and this status is considered variable withsuch parameters as the engine water temperature in the presentembodiment, other parameters such as intake air temperature may beadditionally taken into account in judging the low-temperature state.Alternatively, a temperature estimating device may be provided toestimate the temperature in the following cylinders 2B, 2C or thetemperature in the following cylinders 2B, 2C may be directly orindirectly measured, such that the low-temperature state can be judgedbased on the estimated or measured temperature.

The engine characteristics shown in FIG. 25 need not necessarily be usedin setting the excess-air factor λ of the preceding cylinders 2A, 2D,but other engine characteristics expressed by rightward-descendingcurves may be used. Furthermore, such curves may be more finelysegmented according to the engine speed or other conditions.

While the invention has thus far been described with reference to theseveral embodiments thereof, it is not limited thereto but variousalternatives and changes are possible. Other embodiments of theinvention and variations thereof are described in the following.

(1) Intake and exhaust ports and intercylinder gas channels may bearranged as shown in FIG. 26.

As illustrated in this Figure, the first and fourth cylinders 2A, 2Dwhich are preceding cylinders are each provided with an intake port 11in a left half of the combustion chamber 4 and a first exhaust port 12 aand a second exhaust port 12 b in a right half of the combustion chamber4. Also, the second and third cylinders 2B, 2C which are followingcylinders are each provided with a first intake port 11 a and a secondintake port 11 b in the left half of the combustion chamber 4 and anexhaust port 12 in the right half of the combustion chamber 4. One eachintercylinder gas channel 22 connects the second exhaust ports 12 b ofthe preceding cylinders 2A, 2D to the second intake ports 11 b of thefollowing cylinders 2B, 2C, the intercylinder gas channels 22 runningacross the engine body 1 perpendicular to its cylinder bank. Theconstruction of the engine is otherwise the same as the embodiment shownin FIGS. 1 and 2.

(2) While the flow path switcher is formed by using the valve stopmechanisms 35 in each of the aforementioned embodiments, the flow pathswitcher may be formed by using on-off valves provided in flow channelsas shown in FIG. 27.

As illustrated in this Figure, intake on-off valves 101, 102 areprovided in the branched intake channels 16 individually connected tothe first intake ports 11 a of the second and third cylinders 2B, 2Cwhich are following cylinders, and exhaust on-off valves 103, 104 areprovided in the branched exhaust channels 21 individually connected tothe first exhaust ports 12 a of the first and fourth cylinders 2A, 2Dwhich are preceding cylinders. Further, gas channel on-off valves 105,106 are provided in the intercylinder gas channels 22 between the firstcylinder 2A and the second cylinder 2B, and between the fourth cylinder2D and the third cylinder 2C. Driven by unillustrated actuators, theseon-off valves 101–106 are individually switchable between a state ofopening the respective flow channels (open state) and a state of closingthe respective flow channels (closed state).

The aforementioned on-off valves 101–106 are controlled by anunillustrated controller as follows depending on whether the engineoperating condition falls in the operating range A on the low-load,low-speed side or in the operating range B on the high-load, high-speedside.

-   Operating range A: The intake on-off valves 101, 102 and the exhaust    on-off valves 103, 104 are in the closed state, while the gas    channel on-off valves 105, 106 are in the open state.-   Operating range B: The intake on-off valves 101, 102 and the exhaust    on-off valves 103, 104 are in the open state, while the gas channel    on-off valves 105, 106 are in the closed state.

When the engine operating condition is switched between the operatingranges A and B, the individual on-off valves 101–106 should be switchedbetween their open and closed states during valve switchable periodsshown in FIG. 28. Specifically, if the state of the individual on-offvalves is switched during a period when the exhaust stroke and theintake stroke of a pair of the preceding and following cylindersoverlap, there arises a such problem that the burned gas drawn from thepreceding cylinder mixes with fresh air and they are introduced togetherinto the following cylinder. Thus, the on-off valves 101, 103, 105should be switched within a period excluding the period when the exhauststroke of the first cylinder 2A and the intake stroke of the secondcylinder 2B overlap, and the on-off valves 102, 104, 106 should beswitched within a period excluding the period when the exhaust stroke ofthe fourth cylinder 2D and the intake stroke of the third cylinder 2Coverlap.

The on-off valves 101–106 and the controller for controlling them inthis fashion together constitute the aforementioned flow path switcher.

The intake valves 31, the first and second exhaust valves 32 a, 32 b,the first and second intake valves 31 a, 31 b and the exhaust valves 32provided in the ports of the individual cylinders 2A–2D are caused tocontinually open and close by an unillustrated valve actuatingmechanism. Operation for controlling fuel injection from the individualfuel injectors 9 is the same as that of the foregoing embodiments.

Designated by the numeral 110 in FIG. 27 is a throttle valve provided inthe intake passage 15.

The engine is set to the dual two-cylinder interconnect configuration inthe operating range A and extremely lean mixture combustion is performedin the preceding cylinders 2A, 2D in this embodiment as well. In theoperating range A, the burned gas discharged from the precedingcylinders 2A, 2D is introduced into the respective following cylinders2B, 2C through the intercylinder gas channels 22, combustion in thefollowing cylinders 2B, 2C is made under conditions in which thestoichiometric air-fuel ratio has been produced by a combination of theburned gas of a “lean” air-fuel ratio and newly supplied fuel, and theburned gas discharged from only the following cylinders 2B, 2C is led tothe exhaust passage 20 associated with the three-way catalyst 24. In theoperating range B, on the other hand, the intake ports 11, 11 a and theexhaust ports 12, 12 a of the individual cylinders. 2A–2D workindependently of one another, so that fresh air is introduced throughthe intake passage 15 and the intake ports 11, 11 a into the respectivecylinders 2A–2D and the burned gas discharged through the exhaust ports12, 12 a of the individual cylinders 2A–2D is led to the exhaust passage20. The present embodiment provides the same operational and workingeffects as the earlier-mentioned basic embodiment in this fashion.

According to this embodiment, the construction of the flow path switchercan be made relatively simple. When the engine operating condition isswitched, the on-off valves 101–106 should just be switched during theswitchable periods shown in FIG. 28 and extremely high accuracy is notrequired in their switching timing, so that their control operation iseasy.

(3) Flow channels for the individual cylinders 2A–2D and the flow pathswitcher may be configured as shown in FIG. 29.

As illustrated in this Figure, there are provided intake ports 111 andexhaust ports 112 individually opening to the cylinders 2A–2D of theengine body 1, and an unillustrated valve actuating mechanismcontinually open and close intake valves 113 and exhaust valves 114provided in these ports 111, 112. Branched intake channels 115A–115D areconnected to the intake ports 111 of the cylinders 2A–2D while branchedexhaust channels 116A–116D are connected to the exhaust ports 112 of thecylinders 2A–2D. An intercylinder gas channel 117 is connected between ajoint portion of the branched exhaust channels 116A, 116D of thepreceding cylinders (first and fourth cylinders) 2A, 2D and a jointportion of the branched exhaust channels 116B, 116C of the followingcylinders (second and third cylinders) 2B, 2C, and a first on-off valve118 is provided in the intercylinder gas channel 117.

A joint portion of the branched intake channels 115A, 115D of thepreceding cylinders 2A, 2D is always connected to an upstream portion ofthe intake passage 15, and a second on-off valve 119 is provided in aconnecting portion between a joint portion of the branched intakechannels 115B, 115C of the following cylinders 2B, 2C and the upstreamportion of the intake passage 15 for opening and closing this connectingportion. On the other hand, the joint portion of the branched exhaustchannels 116B, 116C of the following cylinders 2B, 2C is alwaysconnected to a downstream portion the exhaust passage 20, and a thirdon-off valve 120 is provided in a connecting portion between the jointportion of the branched exhaust channels 116A, 116D of the precedingcylinders 2A, 2D and the downstream portion the exhaust passage 20 foropening and closing this connecting portion.

The aforementioned on-off valves 118–120 are controlled by anunillustrated controller as follows depending on whether the engineoperating condition falls in the operating range A on the low-load,low-speed side or in the operating range B on the high-load, high-speedside.

-   Operating range A: The first on-off valve 118 is in the open state,    while the second and third on-off valves 119, 120 are in the closed    state.-   Operating range B: The first on-off valve 118 is in the closed    state, while the second and third on-off valves 119, 120 are in the    open state.

The on-off valves 118–120 and the controller for controlling them inthis fashion together constitute the aforementioned flow path switcher.Operation for controlling fuel injection from the individual fuelinjectors 9 is the same as that of the foregoing embodiments.

The engine is set to the dual two-cylinder interconnect configuration inthe operating range A, in which the burned gas discharged from thepreceding cylinders 2A, 2D is introduced directly into the followingcylinders 2B, 2C through the intercylinder gas channel 117 between twocylinders of which intake and exhaust strokes overlap, and the burnedgas discharged from only the following cylinders 2B, 2C is led to theexhaust passage 20 associated with the three-way catalyst 24 in thisembodiment as well. In the operating range B, on the other hand, theintake ports 111 and the exhaust ports 112 of the individual cylinders2A–2D work independently of one another, so that fresh air is introducedthrough the intake passage 15 and the intake ports 111 into therespective cylinders 2A–2D and the burned gas discharged through theexhaust ports 112 of the individual cylinders 2A–2D is led to theexhaust passage 20.

(4) If ignitability can be maintained even when the fuel is uniformlydispersed in the following cylinders 2B, 2C as described heretofore,fuel injectors provided in the following cylinders 2B, 2C need notnecessarily be of a direct injection type for injecting the fueldirectly into the combustion chambers 4. For example, fuel injectors 9′for injecting the fuel into the following cylinders 2B, 2C may beprovided in the branched intake channels 115B, 115C which constituteintercylinder gas channels as shown in FIG. 30. In this case, uniformcharge combustion is performed in the following cylinders 2B, 2C byinjecting the fuel in the intake stroke from the fuel injectors 9′ toproduce the stoichiometric air-fuel ratio in the following cylinders 2B,2C.

According to this arrangement, heat of exhaust gas introduced from thepreceding cylinders 2A, 2D into the following cylinders 2B, 2C ismoderately dissipated, and the fuel is supplied to a great deal of idealEGR gas in which excess air and burned gas are mixed during a process ofintroducing the gas into the following cylinders 2B, 2C. As a result,evaporation of the fuel and its mixing with the EGR gas are accelerated,and combustibility in the following cylinders 2B, 2C is further improvedwith a great deal of EGR gas introduced thereinto.

(5) Furthermore, a supercharger like a turbocharger 130 shown in FIG.31, for example, may be provided in each of the aforementionedembodiments. Referring to FIG. 31, the turbocharger 130 includes aturbine 131 provided in the exhaust passage 20 and a compressor 132provided in the intake passage 15. The turbine 131 is caused to turn byenergy of the exhaust gas flowing through the exhaust passage 20,whereby the compressor 132 interlocked with the turbine 131 turns andproduces a high intake-air pressure. Designated by the numeral 133 is anintercooler provided in the intake passage 15 downstream of thecompressor 132.

This arrangement makes it possible to achieve a fuel economy improvementeffect by using the aforementioned dual two-cylinder interconnectconfiguration up to relatively high-load regions.

(6) While the gas flow paths are switched by the flow path switcherdepending on whether the engine operating condition falls in theoperating range A on the low-load, low-speed side or in the operatingrange B on the high-load, high-speed side in the foregoing embodiments,the gas flow paths may be connected to form the aforementioned dualtwo-cylinder interconnect configuration in the entire operating rangesof the engine.

(7) The device of this invention is also applicable to multicylinderengines other than the four-cylinder engine. In a six-cylinder engine,for example, the exhaust stroke of one cylinder does not perfectlycoincide in timing with the intake stroke of another cylinder. In such acase, each pair of preceding and following cylinders should be such thatthe exhaust stroke of one cylinder precedes and coincides in part withthe intake stroke of the other cylinder.

(8) The construction of the foregoing embodiments may be modified suchthat EGR operation is performed on the preceding cylinders 2A, 2D only.Since the amount of NOx generated in the preceding cylinders 2A, 2D isreduced and the burned gas introduced from the preceding cylinders 2A,2D into the following cylinders 2B, 2C serves to reduce the amount ofNOx generated therein in the same way that the EGR operation works onthe preceding cylinders 2A, 2D, it is possible to effectively reduce NOxemissions.

INDUSTRIAL APPLICABILITY

As thus far described, the control device of the invention serves tosignificantly improve fuel economy due to an improvement in thermalefficiency and a reduction in pumping loss in the preceding cylindersachieved by lean burn operation therein and yet realize sufficientemission-cleaning performance by use of a three-way catalyst alone.Accordingly, the provision of a lean NOx catalyst becomes unnecessary,resulting in a cost reduction, and it becomes unnecessary to temporarilydecrease the air-fuel ratio (to produce a richer mixture) during leanburn operation, making is possible to avoid a decrease in fuel economyimprovement effect due to temporary enrichment of the mixture. Inaddition, the control device of the invention serves to get rid of theearlier-mentioned sulfur-poisoning problem of the lean NOx catalyst.

Furthermore, high-temperature burned gas discharged from the precedingcylinders is introduced into the following cylinders throughintercylinder gas channels. The provision of the intercylinder gaschannels serves to regulate gas temperature by heat dissipation alongtheir length and sufficiently mix the burned gas with excess air,thereby offering improved combustibility in the following cylinders.

1. A control device for a multicylinder spark-ignition engine of whichindividual cylinders go through successive cycles of intake,compression. expansion and exhaust strokes with specific phase delays,in which fresh air and gas flow paths are connected to form atwo-cylinder interconnect configuration at least in a low-load,low-speed operating range such that, in a pair of preceding andfollowing cylinders whose exhaust and intake strokes overlap each other,burned gas discharged from the preceding cylinder which is currently inthe exhaust stroke is introduced directly into the following cylinderwhich is currently in the intake stroke through an intercylinder gaschannel and gas discharged from only the following cylinder is led to anexhaust passage provided with a three-way catalyst, said control devicecomprising: a controller for controlling fuel supply to the individualcylinders in such a manner that combustion in the preceding cylinder ismade under lean mixture conditions at an air-fuel ratio larger than thestoichiometric air-fuel ratio by a specific amount, fuel for thefollowing cylinder is supplied in an amount corresponding to the burnedgas which was generated by combustion in the preceding cylinder, and awhole amount of which is supplied to the following cylinder andcombustion in the following cylinder is made under conditions of thestoichiometric air-fuel ratio when the engine is in the two-cylinderinterconnect configuration; and a flow path switcher for switching thefresh air and gas flow paths to form an independent cylinderconfiguration in a high-load, high-speed operating range, in whichintake ports and exhaust ports of the individual cylinders workindependently of one another such that fresh air is introduced throughan intake passage into the intake ports of the individual cylinders andexhaust gas discharged through the exhaust ports of the individualcylinders is led to the exhaust passage; wherein said controller makesthe air-fuel ratio in the individual cylinders equal to or smaller thanthe stoichiometric air-fuel ratio in the high-load, high-speed operatingrange.
 2. The control device for the spark-ignition engine according toclaim 1, wherein said preceding cylinder is provided with an intake portconnected to the intake passage, a first exhaust port connected to theexhaust passage and a second exhaust port connected to the intercylindergas channel, while said following cylinder is provided with a firstintake port connected to the intake passage, a second intake portconnected to the intercylinder gas channel and an exhaust port connectedto the exhaust passage, said flow path switcher including: a valve stopmechanism which individually switches first and second exhaust valvesfor opening and closing the first and second exhaust ports of thepreceding cylinder as well as first and second intake valves for openingand closing the first and second intake ports of the following cylinderbetween activated and deactivated states; and a valve stop mechanismcontroller which sets the first exhaust valve and the first intake valveto the deactivated state and the second exhaust valve and the secondintake valve to the activated state in the low-speed, low-load operatingrange, and sets the first exhaust valve and the first intake valve tothe activated state and the second exhaust valve and the second intakevalve to the deactivated state in the high-load, high-speed operatingrange.
 3. The control device for the spark-ignition engine according toclaim 1, wherein the engine is made switchable between special operationmode in which combustion is made in said two-cylinder interconnectconfiguration and normal operation mode in which combustion is made withthe intake ports and the exhaust ports of the individual cylindersworking independently of one another according to engine operatingconditions, said control device further comprising: an intake airpulsation detector for detecting intake air pulsations; wherein saidcontrol device judges at switching of the engine operation mode whetherthe fresh air and gas flow paths have been switched by the flow pathswitcher with reference to a sensing signal output from the intake airpulsation detector and performs air-fuel ratio control operationcorresponding to the operation mode selected after switching of the flowpaths following a point in time when the switching of the flow paths hasbeen detected.
 4. The control device for the spark-ignition engineaccording to claim 3, wherein the control device judges that theswitching of the flow paths has been completed at a point in time when asudden change in the period of intake air pulsations is verified withreference to a sensing signal output from an intake air quantitydetector for detecting the amount of intake air.
 5. The control devicefor the spark-ignition engine according to claim 4, wherein the controldevice judges at the switching of the engine operation mode that theswitching of the flow paths from the two-cylinder interconnectconfiguration to the independent cylinder configuration has been donewhen it is verified that the period of intake air pulsations has becomeshorter.
 6. The control device for the spark-ignition engine accordingto claim 3, wherein the intake and exhaust flow paths are changed byvarying the amount of valve lift determined by a valve actuatingmechanism provided to each cylinder.
 7. The control device for thespark-ignition engine according to claim 3, wherein multiple pairs ofthe preceding and following cylinders of which intake and exhauststrokes overlap with each other are provided, and the control deviceperforms the air-fuel ratio control operation corresponding to theoperation mode selected after the switching of the flow paths in all thepairs of the preceding and following cylinders following a point in timewhen the switching of the flow paths has been first verified in one ofthe multiple pairs of the multiple pairs of the preceding and followingcylinders.
 8. The control device for the spark-ignition engine accordingto claim 3, wherein the control device judges that the flow paths havebeen switched at a point in time when a switching signal has been outputto the flow path switcher following a change in the engine operatingconditions and the occurrence of a change in intake air pulsations hasbeen verified with reference to the sensing signal output from theintake air pulsation detector.
 9. The control device for thespark-ignition engine according to claim 8, wherein the control devicebegins preparation for executing the air-fuel ratio control operationafter the switching of the engine operation mode at a point in time whenit is verified that the flow path switching signal has been output tothe flow path switcher.
 10. A control device for a multicylinderspark-ignition engine of which individual cylinders go throughsuccessive cycles of intake, compression, expansion and exhaust strokeswith specific phase delays, in which fresh air and gas flow paths areconnected to form a two-cylinder interconnect configuration at least ina low-load, low-speed operating range such that, in a pair of precedingand following cylinders whose exhaust and intake strokes overlap eachother, burned gas discharged from the preceding cylinder which iscurrently in the exhaust stroke is introduced directly into thefollowing cylinder which is currently in the intake stroke through anintercylinder gas channel and gas discharged from only the followingcylinder is led to an exhaust passage provided with a three-waycatalyst, said control device comprising: a controller for controllingfuel supply to the individual cylinders in such a manner that combustionin the preceding cylinder is made under lean mixture conditions at anair-fuel ratio larger than the stoichiometric air-fuel ratio by aspecific amount, fuel for the following cylinder is supplied in anamount corresponding to the burned gas which was generated by combustionin the preceding cylinder, and a whole amount of which is supplied tothe following cylinder and combustion in the following cylinder is madeunder conditions of the stoichiometric air-fuel ratio when the engine isin the two-cylinder interconnect configuration; an exhaust gasconcentration detector disposed in the exhaust passage provided with thethree-way catalyst for detecting the stoichiometric air-fuel ratio; andan exhaust gas concentration detector disposed in the intercylinder gaschannel for detecting a lean mixture state; wherein said controllerfeedback-controls the amounts of fuel injected into the individualcylinders based on values detected by the individual exhaust gasconcentration detectors in such a manner that the air-fuel ratio in thepreceding cylinder becomes larger than the stoichiometric air-fuel ratioby a specific amount and the air-fuel ratio in the following cylinderbecomes equal to the stoichiometric air-fuel ratio when the engine is inthe two-cylinder interconnect configuration.
 11. A control device for amulticylinder spark-ignition engine of which individual cylinders gothrough successive cycles of intake, compression, expansion and exhauststrokes with specific phase delays, in which fresh air and gas flowpaths are connected to form a two-cylinder interconnect configuration atleast in a low-load, low-speed operating range such that, in a pair ofpreceding and following cylinders whose exhaust and intake strokesoverlap each other, burned gas discharged from the preceding cylinderwhich is currently in the exhaust stroke is introduced directly into thefollowing cylinder which is currently in the intake stroke through anintercylinder gas channel and gas discharged from only the followingcylinder is led to an exhaust passage provided with a three-waycatalyst, said control device comprising: a controller for controllingfuel supply to the individual cylinders in such a manner that combustionin the preceding cylinder is made under lean mixture conditions at anair-fuel ratio larger than the stoichiometric air-fuel ratio by aspecific amount, fuel for the following cylinder is supplied in anamount corresponding to the burned gas which was generated by combustionin the preceding cylinder, and a whole amount of which is supplied tothe following cylinder and combustion in the following cylinder is madeunder conditions of the stoichiometric air-fuel ratio when the engine isin the two-cylinder interconnect configuration, said engine having afuel injector for injecting fuel directly into the preceding cylinder,wherein said controller causes stratified charge combustion to occur inthe preceding cylinder by injecting the fuel during its compressionstroke from the fuel injector while producing a lean mixture statetherein when the engine is in the two-cylinder interconnectconfiguration.
 12. The control device for the spark-ignition engineaccording to claim 11, said engine further having a fuel injector forinjecting fuel directly into the following cylinder, wherein saidcontroller causes stratified charge combustion to occur in the followingcylinder by injecting at least part of the fuel during its compressionstroke while producing the stoichiometric air-fuel ratio therein whenthe engine is in the two-cylinder interconnect configuration.
 13. Thecontrol device for the spark-ignition engine according to claim 11,wherein said controller causes uniform charge combustion to occur in thefollowing cylinder while producing the stoichiometric air-fuel ratiotherein when the engine is in the two-cylinder interconnectconfiguration.
 14. The control device for the spark-ignition engineaccording to claim 11, said engine further having a fuel injectordisposed in an intake passage of the following cylinder for injectingfuel directly into the following cylinder, the intake passageconstituting the intercylinder gas channel, wherein said controllercauses uniform charge combustion to occur in the following cylinder byinjecting the fuel during its intake stroke while producing thestoichiometric air-fuel ratio therein when the engine is in thetwo-cylinder interconnect configuration.
 15. The control device for thespark-ignition engine according to claim 11, wherein the air-fuel ratioin the preceding cylinder is made approximately equal to twice thestoichiometric air-fuel ratio or larger when the engine is in thetwo-cylinder interconnect configuration.
 16. A control device for amulticylinder spark-ignition engine of which individual cylinders gothrough successive cycles of intake, compression, expansion and exhauststrokes with specific phase delays, in which fresh air and gas flowpaths are connected to form a two-cylinder interconnect configuration atleast in a low-load, low-speed operating range such that, in a pair ofpreceding and following cylinders whose exhaust and intake strokesoverlap each other, burned gas discharged from the preceding cylinderwhich is currently in the exhaust stroke is introduced directly into thefollowing cylinder which is currently in the intake stroke through anintercylinder gas channel and gas discharged from only the followingcylinder is led to an exhaust passage provided with a three-waycatalyst, said control device comprising: a controller for controllingfuel supply to the individual cylinders in such a manner that combustionin the preceding cylinder is made under lean mixture conditions at anair-fuel ratio larger than the stoichiometric air-fuel ratio by aspecific amount, fuel for the following cylinder is supplied in anamount corresponding to the burned gas which was generated by combustionin the preceding cylinder, and a whole amount of which is supplied tothe following cylinder and combustion in the following cylinder is madeunder conditions of the stoichiometric air-fuel ratio when the engine isin the two-cylinder interconnect configuration; wherein the engine iscontrolled in a manner that makes combustion by compressed self-ignitionin the following cylinder at least in part of an operating range inwhich the engine is in the two-cylinder interconnect configuration, andwherein, when the engine is in the operating range in which thetwo-cylinder interconnect configuration is formed and the engine isstill in a specific low-temperature state in which the temperature inthe following cylinder is judged to have not reached a level suitablefor combustion by compressed self-ignition, the air-fuel ratio in thefollowing cylinder is made substantially equal to the stoichiometricair-fuel ratio to make combustion by forced ignition therein, whereasthe air-fuel ratio in the preceding cylinder is made larger than a casewhere combustion is made by compressed self-ignition in the followingcylinder.
 17. The control device for the spark-ignition engine accordingto claim 16, wherein the air-fuel ratio in the preceding cylinder is setsuch that the excess-air factor becomes substantially equal to or largerthan 3 and stratified charge combustion is made in the precedingcylinder when the engine is in said specific low-temperature state. 18.The control device for the spark-ignition engine according to claim 16,wherein the air-fuel ratio in the preceding cylinder is made relativelylarge in a specific low-load region of the operating range in which theengine is in the two-cylinder interconnect configuration compared to ahigher-load region.
 19. The control device for the spark-ignition engineaccording to claim 16, wherein control operation for said specificlow-temperature state is performed when the engine is at or close to itsidling speed.
 20. A control device for a four-cycle multicylinderspark-ignition engine of which individual cylinders provided with sparkplugs go through successive cycles of intake, compression, expansion andexhaust strokes with specific phase delays, said engine having anintercylinder gas channel interconnecting a pair of preceding andfollowing cylinders whose exhaust and intake strokes overlap each otherfor introducing burned gas discharged from the preceding cylinder whichis currently in the exhaust stroke into the following cylinder which iscurrently in the intake stroke; wherein said preceding cylinder isprovided with an intake port connected to an intake passage, a firstexhaust port connected to an exhaust passage and a second exhaust portconnected to the intercylinder gas channel, while said followingcylinder is provided with a first intake port connected to the intakepassage, a second intake port connected to the intercylinder gas channeland an exhaust port connected to the exhaust passage; wherein first andsecond exhaust valves for opening and closing the first and secondexhaust ports of the preceding cylinder as well as first and secondintake valves for opening and closing the first and second intake portsof the following cylinder are made individually switchable betweenactivated and deactivated states; wherein a two-cylinder interconnectconfiguration is formed in a low-load, low-speed operating range suchthat burned gas discharged from the preceding cylinder which iscurrently in the exhaust stroke is introduced into the followingcylinder which is currently in the intake stroke by bringing the secondexhaust valve and the second intake valve to the activated state;wherein exhaust gas discharged from the exhaust port of the followingcylinder is caused to go through a thee-way catalyst provided in theexhaust passage when the engine is in the two-cylinder interconnectconfiguration; said control device comprising: a controller forcontrolling fuel supply to the individual cylinders in such a mannerthat combustion in the preceding cylinder is made under lean mixtureconditions at an air-fuel ratio larger than the stoichiometric air-fuelratio by a specific amount, fuel for the following cylinder is suppliedto the burned gas generated by combustion in the preceding cylinder, andcombustion in the following cylinder is made under conditions of thestoichiometric air-fuel ratio when the engine is in the two-cylinderinterconnect configuration.
 21. The control device for thespark-ignition engine according to claim 20, wherein the engine is madeswitchable between special operation mode in which combustion is made insaid two-cylinder interconnect configuration and normal operation modein which combustion is made under conditions in which intake and exhaustoperations of the individual cylinders are performed independently ofone another with the first exhaust valve and the first intake valve setto the activated state and the second exhaust valve and the secondintake valve set to the deactivated state according to engine operatingconditions, said control device further comprising: an intake airpulsation detector for detecting intake air pulsations; wherein saidcontrol device judges at switching of the engine operation mode whetherfresh air and gas flow paths have been switched by a flow path switcherwith reference to a sensing signal output from the intake air pulsationdetector and performs air-fuel ratio control operation corresponding tothe operation mode selected after switching of the flow paths followinga point in time when the switching of the flow paths has been detected.22. The control device for the spark-ignition engine according to claim20, said controller including: a total fuel injection quantitycalculator for calculating the sum of the amounts of fuel to be injectedinto the preceding cylinder and the following cylinder based on theamount of intake air introduced into the preceding cylinder in such amanner that combustion in the preceding cylinder is made under said leanmixture conditions at the air-fuel ratio larger than the stoichiometricair-fuel ratio by the specific amount and combustion in the followingcylinder is made under said conditions of the stoichiometric air-fuelratio when the engine is in the two-cylinder interconnect configuration;a ratio setter for setting a ratio of the air-fuel ratio for thepreceding cylinder to the air-fuel ratio for the following cylinderaccording to engine operating conditions in such a manner that a balanceis achieved between a torque generated by the preceding cylinder and atorque generated by the following cylinder when the engine is in thetwo-cylinder interconnect configuration; and a final fuel injectionquantity calculator for calculating final amounts of fuel to be injectedinto the preceding cylinder and the following cylinder based on theratio of the air-fuel ratios set by the ratio setter and the sum of theamounts of fuel to be injected calculated by the total fuel injectionquantity calculator.
 23. The control device for the spark-ignitionengine according to claim 20, wherein the engine is so controlled as tomake combustion by compressed self-ignition in the following cylinder atleast in part of an operating range, and wherein when the engine is inthe operating range in which the two-cylinder interconnect configurationis formed and the engine is still in a specific low-temperature state inwhich the temperature in the following cylinder is judged to have notreached a level suitable for combustion by compressed self-ignition, theair-fuel ratio in the following cylinder is made substantially equal tothe stoichiometric air-fuel ratio to make combustion by forced ignitiontherein, whereas the air-fuel ratio in the preceding cylinder is madelarger than a case where combustion is made by compressed self-ignitionin the following cylinder.
 24. A control device for a multicylinderspark-ignition engine of which individual cylinders go throughsuccessive cycles of intake, compression, expansion and exhaust strokeswith specific phase delays, in which fresh air and gas flow paths areconnected to form a two-cylinder interconnect configuration at least ina low-load, low-speed operating range such that, in a pair of precedingand following cylinders whose exhaust and intake strokes overlap eachother, burned gas discharged from the preceding cylinder which iscurrently in the exhaust stroke is introduced directly into thefollowing cylinder which is currently in the intake stroke through anintercylinder gas channel and gas discharged from only the followingcylinder is led to an exhaust passage provided with a three-waycatalyst, said control device comprising: a control unit for controllingthe engine; wherein said control unit controls fuel supply to theindividual cylinders in such a manner that combustion in the precedingcylinder is made under lean mixture conditions at an air-fuel ratiolarger than the stoichiometric air-fuel ratio by a specific amount, fuelfor the following cylinder is supplied in an amount corresponding to theburned gas which was generated by combustion in the preceding cylinder,and a whole amount of which is supplied to the following cylinder andcombustion in the following cylinder is made under conditions of thestoichiometric air-fuel ratio when the engine is in the two-cylinderinterconnect configuration, and wherein the engine is made switchablebetween special operation mode in which combustion is made in saidtwo-cylinder interconnect configuration and normal operation mode inwhich combustion is made with intake ports and exhaust ports of theindividual cylinders working independently of one another according toengine operating conditions; and an airflow sensor for detecting intakeair pulsations; wherein said control unit judges at switching of theengine operation mode whether the fresh air and gas flow paths have beenswitched by a flow path switcher with reference to a sensing signaloutput from the airflow sensor and performs air-fuel ratio controloperation corresponding to the operation mode selected after switchingof the flow paths following a point in time when the switching of theflow paths has been detected.
 25. The control device for thespark-ignition engine according to claim 24, wherein said control unitfor controlling the engine; calculates the sum of the amounts of fuel tobe injected into the preceding cylinder and the following cylinder basedon the amount of intake air introduced into the preceding cylinder insuch a manner that combustion in the preceding cylinder is made undersaid lean mixture conditions at the air-fuel ratio larger than thestoichiometric air-fuel ratio by the specific amount and combustion inthe following cylinder is made under said conditions of thestoichiometric air-fuel ratio when the engine is in the two-cylinderinterconnect configuration; sets a ratio of the air-fuel ratio for thepreceding cylinder to the air-fuel ratio for the following cylinderaccording to engine operating conditions in such a manner that a balanceis achieved between a torque generated by the preceding cylinder and atorque generated by the following cylinder when the engine is in thetwo-cylinder interconnect configuration; and calculates final amounts offuel to be injected into the preceding cylinder and the followingcylinder based on said ratio of the air-fuel ratios and said sum of theamounts of fuel to be injected.
 26. A control device for a multicylinderspark-ignition engine of which individual cylinders go throughsuccessive cycles of intake, compression, expansion and exhaust strokeswith specific phase delays, in which fresh air and gas flow paths areconnected to form a two-cylinder interconnect configuration at least ina low-load, low-speed operating range such that, in a pair of precedingand following cylinders whose exhaust and intake strokes overlap eachother, burned gas discharged from the preceding cylinder which iscurrently in the exhaust stroke is introduced directly into thefollowing cylinder which is currently in the intake stroke through anintercylinder gas channel and gas discharged from only the followingcylinder is led to an exhaust passage provided with a three-waycatalyst, said control device comprising: a controller for controllingfuel supply to the individual cylinders in such a manner that combustionin the preceding cylinder is made under lean mixture conditions at anair-fuel ratio larger than the stoichiometric air-fuel ratio by aspecific amount, fuel for the following cylinder is supplied in anamount corresponding to the burned gas which was generated by combustionin the preceding cylinder, and a whole amount of which is supplied tothe following cylinder and combustion in the following cylinder is madeunder conditions of the stoichiometric air-fuel ratio when the engine isin the two-cylinder interconnect configuration; wherein said controlunit for controlling the engine controls the engine in a manner thatmakes combustion by compressed self-ignition in the following cylinderat least in part of an operating range in which the engine is in thetwo-cylinder interconnect configuration, and wherein, when the engine isin the operating range in which the two-cylinder interconnectconfiguration is formed and the engine is still in a specificlow-temperature state in which the temperature in the following cylinderis judged to have not reached a level suitable for combustion bycompressed self-ignition, the air-fuel ratio in the following cylinderis made substantially equal to the stoichiometric air-fuel ratio to makecombustion by forced ignition therein, whereas the air-fuel ratio in thepreceding cylinder is made larger than a case where combustion is madeby compressed self-ignition in the following cylinder.